Differential planetary gearbox

ABSTRACT

A torque transfer device has plural planets arranged for planetary rotation about one or more sun gears and within one or more ring gears. Each planet includes at least one planetary gear set comprising plural planetary gears connected to rotate together, but having a different diameter to form a differential gear system. To improve load sharing, the plural planetary gears of each planetary gear set may have a different helical angle, the plural planetary gear sets being axially movable with respect to one another. Alternatively or in addition, the planetary gears may be made flexible with respect to radial forces.

BACKGROUND Technical Field

Differential planetary gearboxes.

Description of the Related Art

Planetary gearbox reducers require load balancing between all pinions toensure that all the planets are being used equally. Load balancing isoften accomplished by the use of only three planetary gears so the sungear centers itself by triangulation. If four planets are used, and ifall gears are not perfectly the same size, three gears will take themajority of the load and the fourth gear will carry less than 25% of thetotal load. The more gears that are added, the smaller the planets mustbe, so at a certain number of planets, there is a detriment to maxtorque that can be transferred. As evidence of this, most planetarygearsets in industry have only three planet gears per stage with a smallpercentage having four or five planets per stage.

A differential planetary allows much higher gear ratios with smallerdiameter planet gears. The same load sharing problem exists as with astandard planetary, however, so the use of more than three planets canbe detrimental. As described above, this is because smaller diameterplanets cannot transmit as much load, so if three or four planets takemost of the load, the rest of the planets do not contributeproportionately to the torque transmission and the planets that aredoing the majority of the torque transmission may be too small toprovide benefit as compared to using only three larger diameter planets.

Smaller diameter planets are highly desirable, because they allow alarger center through hole in the gearbox. If load sharing is achievedfor a high number of planets, greater torque transmission is also shownto be possible. In embodiments of a differential gearbox such as thosedisclosed here, FEA analysis has shown that between 12 and 18 smallerplanets provide favorable torque transmission as compared to using only3 larger planets.

What is needed is a way to provide consistent load sharing between theplanets of a differential gearbox so a large number of smaller planetscan be used. There are many applications where high torque output, withminimal weight and strict envelope is required. Other benefits of thedevice will be evident in the description below.

BRIEF SUMMARY

There is provided a torque transfer device having plural planetsarranged for planetary rotation about one or more sun gears and withinone or more ring gears. The plural planets each include a respectivefirst planetary gear set comprising plural planetary gears connected torotate together and having different diameters. A respective firstoutput gear of the one or more sun gears or one or more ring gears isarranged to mesh with a respective planetary gear of each firstplanetary gear set, and a respective first reference gear of the one ormore sun gears or one or more ring gears being arranged to mesh with arespective planetary gear of each first planetary gear set. Load sharingis provided by one or more of A or B or C or D, where:

A is the plural planetary gears of each first planetary gear set havedifferent helical angles and each first planetary gear set is axiallymovable, for example against an elastic element such as a spring withrespect to an axis defined by the one or more sun gears;

B is the plural planets number at least 5, 6, 7, 8, 9, 10, 11, 12, 13,14, 15, 16, 17, 18, 19, or 20 and are formed of a first material withyield strength-to-stiffness ratio greater than 0.10;

C is the plural planetary gears of each first planetary gear set aredefined by a pinion surface and separated by a torsionally flexibleportion of the pinion surface;

D is the plural planets number at least 5, 6, 7, 8, 9, 10, 11, 12, 13,14, 15, 16, 17, 18, 19, or 20 and one or more of the one or more sungears and/or one or more of the one or more ring gears are formed of afirst material with yield strength-to-stiffness ratio greater than 0.10.

In various embodiments, there may be included any one or more of thefollowing features.

Each planet further may also have second planetary gear setcorresponding to and arranged axially symmetrically with respect to thefirst planetary gear set. The second planetary gear set may be arrangedto mesh with a corresponding second output gear of the one or more sungears or one or more ring gears and a corresponding second referencegear of the one or more sun gears or one or more ring gears. The secondplanetary gear set of each planet may have a gear tooth profile axiallysymmetric with respect to a gear tooth profile of the correspondingfirst planetary gear set. Here, the gear tooth profile refers to thethree dimensional shape of the gear teeth on the gears. The first andsecond output gears may be connected via a shim for adjusting therelative axial positioning of the first and second output gears. Thefirst and second reference gears may be connected via a shim foradjusting the relative axial positioning of the first and secondreference gears. Where there is a first input gear as described below,there may also be a second input gear, arranged to mesh with the secondplanetary gear set, and the first input gear and the second input gearmay be connected by a shim for adjusting the relative axial positioningof the first and second input gears. The first reference gear and firstoutput gear may be connected via bearings, the bearings connected to atleast one of the first reference gear and first output gear via a shim.The bearings may be connected to the at least one of the first referencegear and first output gear via plural shims connected to differentbearing races. The second output gear may be the first output gear, andmay have plural gear surfaces or single continuous gear surface. Thesecond reference gear may be the first reference gear, and may haveplural gear surfaces or a single continuous gear surface. The firstreference gear and second reference gear may be ring gears, and thefirst reference gear may be connected to the second reference gear(which rigid connection makes them the same gear, as the notion ofsameness is defined here) via a housing portion extending through acenter hole defined by the one or more sun gears. This housing portionmay define a center bore. There may be gear set spacing springs arrangedto space the respective second planetary gear set of each planet inrelation to the corresponding first planetary gear set. The secondplanetary gear sets may be aligned with the corresponding firstplanetary gear sets by rods, each extending through a respective secondplanetary gear set and the corresponding first planetary gear set. Theremay also be gear set positioning springs on the rods arranged toposition the first planetary gear sets and second planetary gear setsrelative to the rods. The first output gear and the first reference gearmay both be sun gears of the one or more sun gears or both be ring gearsof the one or more ring gears, and connect to different planetary gearsof each first planetary gear set.

Some cases of gear combinations are as follows. One of the pluralplanetary gears of the first planetary gear set may be a spur gear, andanother a helical gear. One of the plural planetary gears of the firstplanetary gear set may be a helical gear, and another a helical gear ofthe same handedness of different magnitude of helix angle. One of theplural planetary gears of the first planetary gear set may be a helicalgear, and another is a helical gear of opposite handedness. Load sharingtechnique A is particularly applicable to the above cases of gearcombinations, and in technique A each first planetary gear set isaxially movable with respect to an axis defined by the one or more sungears. In other cases, one of the plural planetary gears of the firstplanetary gear set may be a helical gear, and another a helical gearwith the same handedness and magnitude of helix angle, or one of theplural planetary gears may be a spur gear, and another also a spur gear.For ease of manufacture, the plural planetary gears of the firstplanetary gear set may have the same number of teeth and correspondingteeth of the plural planetary gears may be circumferentially aligned.Further, the teeth of the plural planetary gears of the first planetarygear set may be connected by a continuous tooth profile fill betweencorresponding teeth. Regardless of the choice of A, B or C for loadsharing, there may be 5 or more, 6 or more, 7 or more, 8 or more, 9 ormore, 10 or more, 11 or more, 12 or more, 13 or more, 14 or more, 15 ormore, 16 or more, 17 or more, 18 or more, 19 or more, or 20 or moreplanets.

The planets of the plural planets may be hollow. The planets of theplural planets may each define a respective axial bore having arespective bore diameter at least ½, ⅔, ⅘, 9/10 or 19/20 of a respectiveouter planet diameter.

For any torque transfer device described above, option B may be chosen,so that the plural planets number at least 5, 6, 7, 8, 9, 10, 11, 12,13, 14, 15, 16, 17, 18, 19, or 20 and are formed of a first materialwith yield strength-to-stiffness ratio greater than 0.10. Particularlywith respect to option B, each planet of the plural planets may have arespective outer pinion portion defining a respective pinion surface.The respective outer pinion portion of each planet may be mounted on arespective hollow tube, the respective outer pinion portions beingformed of the first material. The respective hollow tubes may be formedof a second material stiffer than the first material. The first materialmay be a plain or fiber reinforced polymer resin and the second materialis a metal. The respective pinion surfaces of the plural planets mayeach have an output geared surface arranged to mesh with the output gearand a reference geared surface arranged to mesh with the reference gear,the output geared surface and the reference geared surface separated bya torsionally flexible pinion portion of the respective outer pinionportion. Each torsionally flexible pinion portion may define a recessedportion of the pinion surface. Each torsionally flexible pinion portionmay define axial or radial slots in the pinion surface. The firstmaterial may have a ratio of torsion twist stiffness to bendingstiffness of less than 1.

Option C may also be chosen, including in combination with option B.With option C the plural planetary gears of each first planetary gearset may be defined by a pinion surface and separated by a torsionallyflexible portion of the pinion surface. Particularly with respect tooption C, each torsionally flexible pinion portion may define a recessedportion of the pinion surface. Each torsionally flexible pinion portionmay define axial or radial slots in the pinion surface. Each planet ofthe plural planets may have a respective outer pinion portion definingthe respective pinion surface, the respective outer pinion portion beingmounted on a respective hollow tube. The respective outer pinionportions may be formed of a first material, and the respective hollowtubes are formed of a second material, the second material being stifferthan the first material. The first material may be a plain or fiberreinforced polymer resin and the second material may be a metal. Eachplanet may have a ratio of torsion twist stiffness to bending stiffnessof less than 1.

With respect to any torque transfer device as described above, theplanetary gear sets may be arranged in groups, the planetary gears ofeach group in phase with respect to meshing with other gears, andplanetary gears of different groups not in phase, and the planetarygears of each group may be evenly distributed about the sun gear(s). Theplanets could alternatively be spaced unevenly. Optionally, theplanetary gear sets of each planet may be in phase with one another, sothat the planets as a whole are arranged in such groups.

There may also be a free spinning sun or ring element arranged to engagein traction or geared contact with the planets. The free spinning sun orring element may be a gear of the one or more sun gears or one or morering gears, the free spinning sun or ring element being arranged to meshwith a respective planetary gear of each first planetary gear set. Theremay also be a brake arranged to contact the free spinning sun or ringelement. The free spinning sun or ring element may also have two axiallyseparated contact portions for contacting the planets, the contactportions being oriented to preload the planets depending on a relativeaxial position of the contact portions. The contact portions may bebiased to preload the planets. There may also be an actuation means foradjusting the axial separation of the contact portions.

There may also be a respective first input gear of the one or more sungears or one or more ring gears, the first input gear being arranged tomesh with a respective planetary gear of each first planetary gear set.The first input gear may be connected to an input member, and the firstreference gear connected to a housing member, the input member rotatablyconnected to the housing member through one or more intermediatemembers, the input member rotatably connected to an intermediate memberof the one or more intermediate members through a first set of bearingsand the output member rotatably connected to the intermediate member oranother intermediate member of the one or more intermediate membersthrough a second set of bearings. Two of the first input gear, firstreference gear, and first output gear may be ring gears and one of thefirst input gear, first reference gear, and first output gear may be asun gear. In such a case, the input gear may be a sun gear for a speedreducer. Two of the first input gear, first reference gear, and firstoutput gear may be sun gears and one of the first input gear, firstreference gear, and first output gear may be a ring gear. In such acase, the input gear may be a ring gear for a speed reducer.

There is also provided an actuator combining a torque transfer devicehaving an input gear as described above with a motor connected to drivethe input gear relative to the first reference gear. The actuator mayhave a heat conductive component adjacent to the motor and protrudingthrough a housing to an outer surface of the actuator.

There is also provided an electric device adding, to the torque transferdevice described above, first electromagnetic elements mounted on theplanetary rollers and second electromagnetic elements arranged to act onthe first electromagnetic elements to drive the planetary rollers. Thesecond electromagnetic elements may be connected to the first referencegear. The first electromagnetic elements may be permanent magnets. Thesecond electromagnetic elements may be electromagnets. The secondelectromagnetic elements may have soft magnetic posts or may be aircoils. If air coils, or if using an unusually small soft magnetic post,the stator may use a soft magnetic material backiron without introducingmuch cogging. Thus, there may be a backiron adjacent to the secondelectromagnetic elements. This provides a more efficient air coildesign.

An electric device is also provided without a load sharing scheme asdescribed above. Thus, there is also provided an electric device havingan inner free spinning sun ring, planetary rollers in rolling contactwith the inner free spinning sun ring, an outer fixed ring, an outeroutput ring, the planetary rollers having a first diameter in gearedcontact with the outer fixed ring and a second diameter in gearedcontact with the outer output ring to drive the outer output ringrelative to the outer fixed ring, first electromagnetic elements mountedon the planetary rollers and second electromagnetic elements arranged toact on the first electromagnetic elements to drive the planetaryrollers. The second electromagnetic elements may be connected to theouter fixed ring. There may also be an additional outer fixed ringconnected to the outer fixed ring through a center hole defined by theinner free spinning sun ring. The first electromagnetic elements may bepermanent magnets. The second electromagnetic elements may beelectromagnets. The second electromagnetic elements may be air coils.There may be a backiron adjacent to the second electromagnetic elements.

These and other aspects of the device are set out in the claims.

BRIEF DESCRIPTION OF THE SEVERAL VIEWS OF THE DRAWINGS

Embodiments will now be described with reference to the figures, inwhich like reference characters denote like elements, by way of example,and in which:

FIG. 1 is a side cutaway view of an exemplary actuator including anelectric motor and a differential gearbox that uses planets each havingportions with different helix angles.

FIG. 2 is a closeup side cutaway view of the actuator of FIG. 1.

FIG. 3 is a side cutaway view of the actuator of FIG. 1.

FIG. 4 is an isometric cutaway view of the actuator of FIG. 1.

FIG. 5 is a front view of the actuator of FIG. 1.

FIG. 6 is a side view of the actuator of FIG. 1.

FIG. 7 is a rear view of the actuator of FIG. 1.

FIG. 8 is an isometric view of the actuator of FIG. 1.

FIG. 9 is an exploded view of the actuator of FIG. 1.

FIG. 10 is an exploded view of the actuator of FIG. 1 without housing orinput connector.

FIG. 11 is an exploded view of a gearbox of the actuator of FIG. 1.

FIG. 12 is a side cutaway view of the ring gears of the actuator of FIG.1.

FIG. 13 is an isometric view of the ring gears of the actuator of FIG.1.

FIG. 14 is an exploded view of a planet for the gearbox of the actuatorof FIG. 1.

FIG. 15 is a side section view of the planet of FIG. 14.

FIG. 16 is a side view of the planet of FIG. 14.

FIG. 17 is a side view of the planet of FIG. 14 showing prestressedforces.

FIG. 18 is a side cutaway view of the exemplary actuator of FIG. 1showing forces on a nominal planet.

FIG. 19 is a side cutaway view of the exemplary actuator of FIG. 1showing forces on a planet with one side small.

FIG. 20 is a side cutaway view of the exemplary actuator of FIG. 1showing forces on a planet with both sides small.

FIG. 21 is a front view of the actuator of FIG. 1 with the front outputand housing portions removed.

FIG. 22 is an isometric view of the actuator of FIG. 1 with the frontoutput and housing portions removed.

FIG. 23 is an isometric view of the actuator of FIG. 1 with the frontend springs of the planets also removed

FIG. 24 is a front view of the actuator of FIG. 1 with the front outputand housing portions removed, also schematically showing planetpositioning changes.

FIG. 25 is an isometric view of a planet for the actuator of FIG. 1having aligned teeth.

FIG. 26 is an isometric view of a planet for the actuator of FIG. 1having offset teeth.

FIG. 27 is an isometric cutaway view of an exemplary gearbox.

FIG. 28 is a front view of the gearbox of FIG. 27.

FIG. 29 is a side view of the gearbox of FIG. 27.

FIG. 30 is a rear view of the gearbox of FIG. 27.

FIG. 31 is an isometric view of the gearbox of FIG. 27.

FIG. 32 is a rear view of the gearbox of FIG. 27 with the stationaryring gears and portions of the housing removed.

FIG. 33 is an exploded view of the gearbox of FIG. 27.

FIG. 34 is a side cutaway view of a portion of the gearbox of FIG. 27including a planet gear.

FIG. 35 is a side cutaway view of the gearbox of FIG. 27.

FIG. 36 is a side cutaway view of the gearbox if FIG. 27 with theplanets removed.

FIG. 37 is an isometric cutaway view of the gearbox of FIG. 27.

FIG. 38 is a side view of an exemplary planet for the gearbox of FIG.27.

FIG. 39 is an isometric view of a planet for the gearbox of FIG. 27,formed as one piece of injection molded plastic.

FIG. 40-42 are isometric views of additional embodiments of planetssuitable for the gearbox of FIG. 27.

FIG. 43 is an isometric view of a planet with axial slots and no tube.

FIG. 44 is a cutaway isometric view of the planet of FIG. 43.

FIG. 45 is an isometric view of a planet with axial slots and a centraltube.

FIG. 46 is a cutaway isometric view of the planet of FIG. 45.

FIG. 47 is a cutaway view of the gearbox of FIG. 27 showing axial forceson a planet.

FIG. 48 is an isometric cutaway view of a planet in the gearbox of FIG.27 showing load paths on the planet.

FIG. 49 is a schematic illustration of a gearbox showing outer geardeformation.

FIG. 50 is a schematic illustration of a gearbox showing sun geardeformation.

FIG. 51 is a schematic illustration of a gearbox showing planet geardeformation.

FIG. 52 is a bar chart showing material strength to stiffness to densityratio for selected materials.

FIG. 53 is a bar chart showing material strength to stiffness ratio forselected materials.

FIG. 54 is a cutaway isometric view of another exemplary gearbox.

FIG. 55 is an isometric view of the gearbox of FIG. 54.

FIG. 56 is an axial end view of the gearbox of FIG. 54.

FIG. 57 is an isometric view of an exemplary planet for the gearbox ofFIG. 54.

FIG. 58 is another isometric view of the planet of FIG. 57.

FIG. 59 is an axial end view of the planet of FIG. 57.

FIG. 60 is a side cutaway view of another exemplary embodiment of anactuator.

FIG. 61 is an isometric cutaway view of the actuator of FIG. 60.

FIG. 62 is an isometric cutaway view of the actuator of FIG. 60 withoutmotor input components.

FIG. 63 is a top isometric view of upper motor components for theactuator of FIG. 60.

FIG. 64 is a bottom isometric view of upper motor components for theactuator of FIG. 60.

FIG. 65 is a cutaway isometric view of an upper portion of the actuatorof FIG. 60 showing airflow paths.

FIG. 66 is an isometric view of an exemplary planet for the actuator ofFIG. 60 with continuous features to ease injection molding.

FIG. 67 is an exploded view of an upper half of the actuator of FIG. 60.

FIG. 68 is a side section view of another exemplary planet for theactuator of FIG. 60.

FIG. 69 is a side section view of an expandable sun ring interfacingwith the planet of FIG. 68.

FIG. 70 is a side section view of an adjustable sun ring interfacingwith a planet to provide a safety brake.

FIGS. 71-78 are side section views of different configurations ofgearboxes suitable to act as bearings.

FIG. 79 is an axial end view of another exemplary gearbox.

FIG. 80 is an isometric cutaway view of the gearbox of FIG. 79 showinggearsets.

FIG. 81 is a side view of two planets contacting a sun gear, out ofphase.

FIG. 82 shows a view of a gear tooth with a profile shift.

DETAILED DESCRIPTION

Immaterial modifications may be made to the embodiments described herewithout departing from what is covered by the claims.

Embodiments of the device allow load sharing through the use of one ormore strategies as described below. Other benefits of some embodimentsmay include reducing or preventing backlash, and maintaining appropriateaxial and circumferential location of planets without the use of aplanet carrier.

One strategy is to use planets that have different portions withdifferent helix angles. In this strategy, an input, output and referencegear may each contact the planets. The different gears contacting theplanets may collectively be referred to as i/o gears. For a speedreducer, typically the input will be on one side of the planets (e.g., asun gear) and the output and reference on the other (e.g., ring gears).In this document, a “sun gear” refers to any gear with a radial outersurface meshing with planet gears, and a “ring gear” refers to any gearwith a radially inner surface meshing with planet gears. A “sun ring” isa ring-shaped sun gear, not a ring gear by this definition. The two ofthe input, output or reference gears on one side, in this strategy, havedifferent helix angles, meshing with corresponding helix angles of theplanet gears. This allows load balancing through axial shifting of theplanet gears. To keep overall axial alignment, two axially symmetricsets of gears may be provided, with the planet gears of the two setscombined into single planets with halves connected by axial springs. Theaxial inner ring and/or sun gears of the two sets may also be combined.Embodiments using this strategy are further described in the sectionbelow entitled “DIFFERENTIAL HELIX ANGLE WITH AXIAL SPRING LOCATION.”

Another strategy is to use flexible gears. Gears may be made with theuse of a flexible material such as plastic. A surprising benefit isfound from plastic which might conventionally be expected to have lowertorque to weight than steel. Embodiments using flexible materials arefurther described in the section below entitled “PLASTIC GEARS.”

Flexibility depends not only on the material choice but also on theshape of the gears.

Also described in this document are further applications of thestrategies summarized here. The section entitled “EXAMPLE PLANET DRIVENACTUATOR” provides an additional example of an actuator comprising anelectric motor combined with a speed reducing gearbox.

Additional features are also disclosed that may be combined withembodiments of one or more of the strategies described.

The section entitled “PLANETARY BEARING” describes how a planetarygearbox may also act as a bearing, for example for a motor.

The section entitled “OUT OF PHASE GEARS” describes how differentplanets may mesh with the ring gears at two or more tooth mesh positionsat any given moment to reduce noise and vibration.

Differential Helix Angle with Axial Spring Location

In an embodiment, an electric motor is housed within the gearboxenclosure.

FIGS. 1-24 show an exemplary cylindrical actuator including an electricmotor and a differential gearbox that uses planets each having portionswith different helix angles and elastic elements which bias the pinionportions to a preferred but movable axial position. This allowsconsistent enough load sharing between pinions to take advantage of alarge number of smaller pinions. A simplified electric motor stator androtor is shown inside the housing. An internal electric motor allows thetwo outer ring gears 10, 10 to be attached to housing members 12A, 12Bconnected through an annular housing portion 12C defining the centerbore so the ring gears 10, 10 are held from relative rotation to eachother. Housing members 12A, 12B and 12C comprise portions of housing 12.

A side section view of the exemplary embodiment is shown in FIG. 1.FIGS. 1-3 show cutaway views of this exemplary embodiment and FIGS. 5-8show views of the outside of the actuator.

The motor in this embodiment is configured with an inner stator 22 andouter rotor 13, with the rotor supported by stacked bearing assemblies46 and 48. Bearing assembly 46, as shown here and further describedbelow, comprises a ring 18 that links two sets of bearings 17 and 19.

The outer rotor drives a connecting plate 15, which drives the sun gears14 through a spline fit. Note that instead of using a connecting plate15, it would also be possible to integrate the rotor 13 into the sungears 14. This would enable a more axially compact actuator albeit witha smaller center hole.

The sun gears drive the planets 23 with central straight spur gearteeth. A small amount of backlash is introduced to this interface via atooth offset in order to ensure proper meshing. In the embodiment shownthere are 18 planets.

The planets do not require a carrier as would be found in many planetarygear configurations. Instead, they mesh with an axially outer ring gear10 on the outer helical teeth. Axial location of the planet gears isalso provided within tolerances by the load balancing mechanismdescribed below. Because the axially outer ring gear is stationary, theplanets orbit the sun input as the input rotates. The central spur gearteeth on the planets then mesh with the center ring gears 11. The pitchdiameter of the center spur teeth on the planets is different than thepitch diameter of the helical teeth on the planets, causing adifferential output between the center and axially outer ring gears. Theoutput from the center ring gears 11 then connects to a connecting tube16 with a spline fit and contains a bolt group for fastening to otherparts of the mechanism.

Also shown in FIG. 1 are a heat sink 20 and holes 21 in the housing 12,further described below. FIG. 2 shows a closeup cutaway view of thegears and connecting plate 15 only. Also visible in FIG. 2 is a centralspring 7 connecting two halves A and B of the geared portions of aplanet 23, the halves moveably mounted on a tube 6, and outer springs 9connecting the halves to stops 8 mounted on the tube 6. The function ofthese parts of the planet 23 is further described below in relation toFIGS. 14-20.

FIG. 3 shows the actuator of FIG. 1 with the motor and planets omittedfor clarity. FIG. 4 shows an isometric cutaway view of the actuator ofFIG. 1.

FIGS. 5-8 show external views of the actuator of FIG. 1. FIG. 5 shows afront view of the actuator of FIG. 1. FIG. 6 shows a side view. An outerportion corresponding to one of the axially outer ring gears 10 isvisible, but the axially outer ring gears 10 in this embodiment arefixed to the housing 12 and could alternatively both be enclosed in thehousing. FIG. 7 shows a rear view. 8 shows a front isometric view.

FIGS. 9-11 show exploded views of the actuator of FIG. 1. The actuatormay be connected to an external structure through the housing 12, forexample at a first end 38 of the actuator, and may be connected todriven items through output connector 16, for example at a second end 40of the actuator. The gearbox 42 and motor 44 each take up respectiveannular portions of the actuator in this embodiment.

FIG. 10 shows the actuator without the housing or output connector.Bearing assemblies 46 and 48 are seen, which when the actuator isassembled connect both sides of rotor 13 to the heatsink 20

FIG. 11 shows an exploded view of the gearbox. As can be seen, theaxially inner sun gears 14 in this embodiment are formed as a singlepiece, and the axially inner ring gears 11 are formed as a single piece.

FIG. 12 is a side cutaway view and FIG. 13 is an isometric view of thering gears 10 and 11 of the actuator of FIG. 1, better showing thepatterns of the teeth of these gears.

FIGS. 14-16 are closeup views of planet gears 23 of the actuator of FIG.1, showing a load balancing mechanism.

FIG. 14 is an exploded view of a planet 23, FIG. 15 is a side sectionview, and FIG. 16 is a side view. As shown in these views, an axiallyinner and an outer planet gear 1, 2 are manufactured to act as one piece(gearset A) with symmetric axially inner/outer gears 3, 4 comprisinggearset B. Both gearsets A and B are held in coaxial alignment by acenter tube or rod 6. The term “rod” may encompass the term “tube.” Thefit between the gearsets A and B and the rod 6 is such that axial androtational movement of the gearsets A, B on the tube 6 is possible. Acenter spring 7 between the gearsets A, B, and an outside spring 9between the gearsets A, B and the retaining rings 8 at the ends of theshaft 6 allow axial motion of the gearsets A, B on the tube 6, and alsorotational motion of gearset A relative to the other symmetric set B.The inner and outer gears on each set A and B are created with differenthelical angles (In this example, a helical angle of zero is used for theinner gear 1 but any helical angle can be used as long as the axiallyinner and outer gears have different angles. The gears having the samemagnitude of helix angle with different handedness also provides adifference of helix angle. This helical angle difference must alsocompensate for the different diameter of the inner and outer gears suchthat axial movement of the gearset results in the loading of the innergear teeth on the inner gear ring in the opposite direction of the outergear teeth on the outer gear ring). As a result, any axial movement of agearset A will cause the whole gearset A to rotate due to the largerhelical gear angle on one of the gears 2 as it meshes with the helicalgear teeth on the reference ring gear 10 which is fixed to the housing12. We will refer to a stationary output ring gear 11 in thisdescription for simplicity of explanation) The rotation of the gearset Aduring axial movement of gearset A will cause the axially inner spurgears 1 to rotate relative to the axially inner (output) ring gear 11spur gears (Fig A6) until the spur gears on gear 1 are contacting andtransmitting torque to the axially inner (output) ring gear 11.

The relative forces on and movements of the components of the planets 23in this embodiment are illustrated in FIGS. 17-20. The center spring 7may exert a force 24 and the end springs 9 may exert forces 26 and 28.Optionally, as shown in FIG. 17, the springs may be prestressed so thatin the absence of the external forces, all of these spring forces arenon-zero. This can enable all springs to remain in compression as theforces change, as may be useful for springs using Belleville and wavewashers as shown. For other designs, it could be useful for the springsto be and remain in tension.

An arrow 78 in FIGS. 18-20 indicates direction of motion of the teeth ofthe planet.

Axial movement of the gearset A (or B) is caused by the opposing torquethat is transferred through the pinions when the reference ring gear 10experiences a resisting torque during operation (operation referring tothe torque that is applied by the electric motor rotor 13 to the axiallyinner (input) sun gears 14 through the sun gear flange 15. This torqueapplied to the gearset A will result in a torque being transferredthrough the outer helical gear 2 to the outer helical ring gear 10. Inother embodiments other ring or sun gears could mesh with and transfertorque through the helical gears. The torque applied through thishelical gear mesh will result in an axial force 30 on the gearset A asshown in FIG. 18. This axial force is similar to a nut applying axialforce to a threaded bolt. This axial force 30 results in an axialdisplacement 32 of the gearset A that is opposed by one or more of thesprings. When the axial force of the spring (e.g., the force 24 exertedby center spring 7) equals the axial force experienced by the gearset Aas a result of torque transfer on the helical gear mesh 2, the gearset Areaches an equilibrium where axial motion ceases. Due to the spring rateof the springs 7, 9 (which may be, for example, Belleville washers andwave washers as shown here), each gearset A, B on each of the planets 23will find its own equilibrium where the axial position and resultingrelative rotation results in all of the gearsets in the planet (A and Bon each of the multiple planets 23) transmitting a more consistent loadthan if all the gearsets A were one piece with the symmetric gearsets B.This allows the use of more than 3 pinons with relatively consistentload sharing between all of the pinions. Corresponding forces anddisplacements are also shown for gearset B in FIG. 18. FIG. 18 showsforces and displacements in default conditions with good contact.

FIG. 19 shows forces and displacements if one side is small. Helicalgear 3 is small and does not initially experience an axial force fromthe corresponding gear 10. The forces and displacements on gearset A areinitially the same as those shown in FIG. 18. The displacement ofgearset A increases the center spring force 24 which forces gearset B todisplace in motion 34. This will force gear 4 into contact withcorresponding gear 10 to share some of the load.

FIG. 20 shows the case of both gearsets A and B being small. Neitherhelical gears 2 or 4 would experience an axial force from thecorresponding gears 10 if the gears are in the positions of FIG. 18.Compared with the situation in FIG. 18, the center spring force 24 isless opposed by other axial forces and displaces both gearsets outwardsas shown by arrows 36 to contact the corresponding gears 10 and sharesome of the load.

In this embodiment, torque on the helical gears and corresponding gearsin one rotational direction will result in the pinion gearsets A, B oneach pinion assembly to move inward (toward each other) compressing wavespring 7, which exerts a corresponding outward force 24. When torque isreversed, the gearsets A, B on each planet 23 will move in the oppositeaxial direction, compressing the end springs 9.

Including both gearsets A and B allows balancing of axial forces so thatthe planetary gears do not axially shift out of alignment with the sunand ring gears. The gear system could also work with only one ofgearsets A and B, so long as the axial forces were balanced in anotherway, such as for example by mounting the rod 6 on a planet carrier sothat one end spring 9 could transmit a net axial force from the planetcarrier to the planetary gears to balance the axial force on the gears.The symmetric gearsets A and B also prevent twisting of the planetassemblies by balancing the loading of the planets from end to end.

FIGS. 21-24 show the actuator of FIG. 1 with the front output andhousing portions removed. FIG. 21 is a front view, and FIG. 22 is anisometric view. FIG. 23 is an isometric view with the front end springsof the planets also removed. FIG. 24 schematically shows planetpositioning changes 50 for load sharing. Because all the gears mesh withthe input and output gears, which in the embodiment shown are straight,these relative positioning changes 50 are smaller than shown, but thesmall positioning changes help share the load.

As shown in FIGS. 25-26, the teeth of the gears 1 and 2, and those ofgears 3 and 4, can be aligned (as shown in FIG. 25) or offset (as shownin FIG. 26). As described further below, for some manufacturingtechniques it may be easier to produce gearsets with aligned teeth.

Double Bearing: One of the challenges in selecting bearings for largediameter applications is that the maximum rated rotation speed is oftenlimited due to inconsistencies in the bearing tolerances. In order touse off-the-shelf bearings, but allow for higher rotational speeds,multiple bearings may be stacked concentrically as shown in FIG. 1. Eachbearing would then see a significantly reduced rotation speed.

With each bearing maintaining rolling contact, n bearings would each see1/n of the rotation speed.

This premise consists of two or more concentric bearings with axialconstraint appropriate for the loads applied in that application. Oneembodiment, shown in FIG. 1 consists of a pair of concentric bearings 17and 19, with a supporting ring 18 between them. This ring allows for aprecise fit to each of the bearings, while retaining the bearing in theloaded direction. Axial force can then be transmitted from the rotatingend, through the outer bearing, through the retaining ring, through theinner bearing, and on to the internal shaft.

Aluminum Heat Sink: Because of the fact that the gearbox makes use ofstationary ring gears on either side of the assembly, the motor, inembodiments of the device, is contained within the structure 12connecting the stationary (reference) ring gears 10. This poses apotential problem with heat dissipation within this structure. Incertain applications, the surrounding structure may be constructed of apoor heat conducting material. Heat generated by a motor would then havea highly resistant heat flow path out of the enclosure. By making use ofan aluminum (or other highly heat conductive material) heat sink 20 asshown in FIG. 1, the heat generated by a motor has a significant sink tocollect in, ensuring that the stator does not overheat from periodic andshort temperature spikes, In FIG. 1, the heat sink is exposed to theexternal support structure through holes 21 in the enclosure, allowingheat energy to be efficiently dissipated by contact with another heatconducting body in the structure, or by convection cooling, without theheat needing to pass through the enclosure material directly.

Other arrangements of the input, output and reference could also beused. In general, for the gear system to act as a large ratio gearreducer (amplifying torque), as in the embodiment shown in the figures,the output and reference should be both radially inner or radially outerwith respect to the planet arrangement, and the input should be theopposite, connected to either of the gears the output or reference isconnected to, or, in principle, to still another gear. For the system toact as a large ratio gear increaser (reducing torque), the input andreference should be both radially inner or radially outer with respectto the planet arrangement, and the output should be the opposite,connected to either of the gears the input or reference is connected to,or, in principle, to still another gear. For the system to act as asmall ratio gear increaser or reducer, the input and the output shouldbe both radially inner or radially outer with respect to the planetarrangement, and the reference should be the opposite, connected toeither of the gears the input or output is connected to, or, inprinciple, to still another gear.

Where a planet gear connected to one of the input, output or referencein one of the radially inward or outward directions, but not connectedto any of those in the other of the radially inward or outwarddirections, a floating gear can be added if desired.

In an embodiment, a high torque LiveDrive™ electric motor may be used,such as disclosed in U.S. Pat. No. 9,755,463, the content of which ishereby incorporated by reference in its entirety.

Differential gearboxes are typically lower efficiency because the fulloutput torque is also meshing at high speed resulting in a highpercentage of work required to overcome this friction. But with thehigher motor torque at lower speed of the LiveDrive™, a significantlylower gear ratio can be used, thus increasing efficiency. With 18planets and a mechanism to ensure proper load sharing, the contact ratiois 6× that of a traditional planetary gearbox, thus increasing torquecapabilities significantly. The low-ratio allows the input rotor to runslower for a given output speed. This allows for the use of low-profilebearings, which when radially stacked as described above, results inlower total bearing weight. Low maintenance: Very few moving parts.Frequency of maintenance activities and overall downtime are expected tobe reduced.

The actuator may be used for high torque applications where a hydraulicactuator might otherwise be used. Advantages over hydraulics include thefollowing. Highly reliable: damage to one power line does not affectmultiple actuators. Higher MTBF with electrical actuation. Easy tomonitor: only the actuator needs monitoring. Controllability: highlyresponsive and precise due to low-ratio gear-reduction. Electricactuation allows more sophisticated control algorithms. Environmentallysafe: no possibility of leakage or dangerous emissions.

Plastic Gears

The use of plastic gears in a planetary gear box is generally expectedto provide lower torque and lower torque to weight than a steel gearboxof the same size and geometry. In one simple example, a conventionalsteel planetary gearbox using high strength steel for all the gears andhousing and using three large planets would provide approximately threetimes the torque-to-weight of the same gearbox made from carbonreinforced PEEK (which is a very strong injection moldable plastic).

When configured as shown here, the use of plastic gears is believed tohave the potential of providing a surprising result, which is toapproach or even exceed the torque to weight of an equivalentconstruction gearbox made from high strength steel.

FIGS. 27-37 show a non-limiting example of this construction. As shownin FIG. 27, an array of 12 planets 123 are constructed to drive anoutput ring gear 111 relative to stationary ring gears 110 with a 15:1differential ratio when driven with a sun gear 114. Note that forclarity, the frame of reference describing the gears is relative to theaxial direction of the gearbox, with the center gear teeth closer to theaxial center of the gearbox and the outer gear teeth further from theaxial center of the gearbox as labeled in FIG. 27. The sun gear 114 hasteeth 154 that can be designed to mesh with either the center planetgear teeth 101 (which are preferably larger in pitch diameter) or theouter planet gear teeth 102 (which are preferably smaller pitchdiameter). A motor rotor (not shown) may be positioned on the ID of thesun gear 114. In the embodiment shown, the sun gear has an exposedgrooved portion 156 that is designed for input by hand as ademonstration. In another embodiment, the grooves could be replaced bygear teeth so that the grooved portion 156 may mesh with an input gear(not shown). If the input gear is smaller than the sun gear 114 at theexposed gear portion 156 a further gear reduction may be obtained. Adirect input could be provided here with limited range of motion, oralternatively, the housing 112 could be enlarged such that it surroundsa motor, which allows for continuous input.

The sun gear 114 is supported in this embodiment from the housing 112 bya bearing 158 on a bearing sleeve 160.

The center planet gear teeth mesh with the center ring gears 111 whichin an embodiment are the output of the gearbox. The outer gear teeth 102on the planets mesh with the two outer ring gears 110 which, in thisexemplary embodiment, are attached to ground via housing 112. Ballbearings 117 may support the output ring 111 for rotation relative tothe stationary rings 110. A bearing retaining ring 152 in thisembodiment is attached to the output.

Rotation of the sun gear 114 causes the planets 123 to rotate and toorbit around the ID of the ring gears 110 and 111. As a result of thedifferent ratios between the inner planets and the inner ring gear 111,and the outer planet gears and the outer ring gear 110, a differentialreduction is provided which, in this case, equals approximately 15:1with the following gear tooth numbers:

TABLE 1 Gear Number of Teeth Pitch Diameter (mm) Sun gear 84 82.63 Innerring gear (Output) 132 129.85 Outer ring gear 144 127.5 (Stationary)Inner planet gear 24 23.61 Outer planet gear 24 21.25

FIGS. 28-32 show further views of the gearbox of FIG. 27. FIG. 28 is afront view, FIG. 29 is a side view, and FIG. 30 is a rear view. As shownin FIG. 30, the housing 112 may have static mounting features/keyways162 on an inner diameter of housing 112 to better enable the gearbox tobe mounted to another object. FIG. 31 is an isometric view. FIG. 32 is arear view with the stationary ring gears, and portions of the housingthat would obscure the planet gears 123, removed.

FIG. 33 is an exploded view of the gearbox of FIG. 27. As seen in FIG.33, an inner shim 164 allows axial adjustment of the bearing sleeverelative to the housing 112, thus allowing axial adjustment of the sungear 114, and an outer shim 166 separates the two output gears 111 fromeach other, allowing axial adjustment of the output gears 111. Uses ofsuch axial adjustments are described further below in relation to FIGS.54-56. FIG. 33 also shows the teeth 174 of the output gears 111 and theteeth 176 of the stationary gears 110.

FIG. 34 shows a side cutaway view of a portion of the gearbox of FIG. 27including a planet gear 123. FIG. 35 shows a side cutaway view of thewhole gearbox of FIG. 27. FIG. 36 shows a side cutaway view with theplanets removed. This allows the teeth 174 of the output gears to beseen. FIG. 37 shows an isometric cutaway view.

In embodiments of the device, at least one outer planet gear on eachplanet gear assembly must be rotationally fixed to an inner planet gearso torque can be transmitted from a fixed ring gear to the inner ringgear output. The inner planet gear on each planet may be a single gearsuch as a spur gear or two symmetrical helical gears as shown in FIG.38. Note that, in flexibility-based embodiments, it is not necessary forthe inner planet gear teeth 101 and outer planet gear teeth 102 to havedifferent helical angles. FIG. 39 shows a preferred configuration of aplastic planet gear for this exemplary device made of plastic where theouter planet gear teeth and inner planet gear teeth are made of onepiece of injection molded plastic. A continuous tooth profile fill 168connects the inner teeth 101 to the outer teeth 102. The helix angle onthe inner and outer is similar or the same which allows the gears to bepulled out of a two part mold. One half of the mold comprising thenegative of the clockwise helix teeth and the other half of the moldcomprising the negative of the counter clockwise helix teeth.

In a preferred embodiment, the number of teeth on the inner planet gearsand outer planet gears is the same to allow the part to be removed fromthe mold, or to allow gear tooth cutters to shape the teeth withoutinterference with either the inner or outer teeth.

Load sharing in this embodiment is accomplished with plastic gears (andpossibly a plastic housing) as follows:

The use of steel gears in this exemplary embodiment can be used but areproblematic because steel gears are very rigid and are subject tovariations in manufacturing tolerances, especially in a low costactuator. As a result, only 3-5 of the 12 planet gears would be expectedto transmit a higher percentage of the torque if the gearbox is made ofsteel.

By using plastic for the one or two or all of the planet gears and/orthe ring gears, and/or the sun gears (and possibly the housings) a moreflexible assembly is created. As an example, carbon fiber PEEK may havea tensile modulus of 3200 ksi (22 gpa) while a high strength steel suchas maraging steel would have a tensile modulus of 27600 ksi (190 gpa).Maraging steel is stronger than carbon fiber PEEK. In a conventionalplanetary gearbox with three large planets per stage, this would give asteel gearbox greater torque-to-weight compared to a carbon fiber PEEKgearbox of the same design. In the embodiments shown, plastic planetgears are used. However, it would also be possible to use plastic ringand/or sun gears in combination with metal planets. This would haveperformance and wear life benefits, and would still provide adequateload sharing, for example via the shape changes shown in FIGS. 49 and50, if the consistency of the metal parts is high.

As the number of planets is increased, however, the load sharing of asteel gearbox, with its high tensile modulus (high stiffness) willresult in a reduction in the load sharing consistency (as a result ofslight variations in the manufacturing tolerances). By contrast, acarbon fiber PEEK gearbox may have approximately 6× lower stiffness fromthe components which allows the teeth to flex much more than the steelgears. This flexibility would, in a conventional three-pinon-per-stageplanetary gearbox, result in 6X the flexibility for a given size, and asmuch as an estimated 5× the flexibility for the same torque-to-weight.This would be seen as a detrimental combination of effects and wouldsteer a designer away from using plastic gears in a gearbox where hightorque is required from a given size or given weight.

Proposed here is a way to provide torque-to-weight and torque to sizefrom a plastic gearbox that is higher than would be expected, bycombining plastic gears of a certain range of stiffness-to-strength witha gearbox design having a high number of planets such as disclosed here.

The stiffness of steel or other metal, is ordinarily seen as a benefitin terms of creating a gearbox with high stiffness but in a device likeembodiments shown here with many pinions, the stiffness of the steelgears can actually be a detriment because it can reduce the consistencyof load sharing.

The higher flexibility of plastic is believed to be beneficial to thetorque to weight of embodiments of the present device because at acertain number of pinions, the high stiffness of the steel gears becomesdetrimental to load sharing while the flexibility of the plastic gearsallows load sharing to be more consistent above this number of pinions.The result is believed to be a range of high pinion number gearboxgeometry that provides better performance in certain regards such astorque to weight when the geometry of embodiments of the device arecombined with the use of plastic gears rather than steel gears, or evena combination of steel and plastic such as but not limited to steelpinons and plastic ring gears and sun gears.

Above some number of pinions and below a certain strength to stiffnessratio (more flexible is better for load sharing), the increased loadsharing which results from the flexibility of the plastic (or mechanicaltorsion flex member between the inner and outer pinion gears (which alsoapplies to metal gears) it will be possible to achieve higher torquewith a weaker material. The looser the manufacturing tolerances in theconstruction of the gears, the more variability in the gears and themore benefit will be obtained from torsional flexibility of the pinions.

The result is the potential for a very low cost gearbox by virtue ofmaking all or part of it injection moldable combined with much betterperformance in terms of torque to weight than would be expected from aplastic gearbox and possibly even similar or better torque to weightthan a steel gearbox of the same design made with reasonablemanufacturing tolerances.

As shown in FIG. 40, another embodiment of a pinion, suitable for use ina gearbox of the embodiment of FIG. 27, has outer and inner gears whichare all one piece and made of plastic (although other materials can beused with various effects). An optional cylindrical section 170 islocated between the inner and outer gears 101 and 102 (and possiblybetween the two inner gears as shown in FIG. 41). A steel or aluminum orother material (such as possibly a thick walled plastic or thin-walledmetal) bar or tube 106 provides bending stiffness so lengthwise bendingof the pinion is reduced during torque transmission. The plastic gear iscreated with an ID that is slightly larger than the OD of the tube orbar such that the plastic gear is able to rotate freely on the tube 106.The plastic gear may be bounded axially on the tube 106 by stops 108,for example lips as shown here. The tube can also be omitted as shown inFIG. 42. The optional cylindrical sections between the inner and outergears is thin enough that the opposing torque direction on the inner andouter gears of a pinion will result in a small amount of torsional twistof the cylindrical sections. This torsional stiffness can be decreasedby increasing the length of the cylindrical sections or by decreasingthe thickness of these cylindrical sections and/or by creating slots 172in the cylindrical sections. The slots may be for example axial orhelical. FIGS. 43 and 44 show a version with axial slots 172, with noinner tube, and FIGS. 45 and 46 show a version with axial slots 172 andan inner tube 106. Decreasing the torsional stiffness of the cylindricalsections is considered to be beneficial in this design because thetwisting of cylinder sections allows the inner and outer gears on thepinion to rotate slightly relative to each other. As a result,manufacturing intolerance can be compensated for by this relativerotation so a high number of pinions can be used while still achieving ahigh consistency of load sharing.

It should be noted that plastic gears are used as an example here, butmetal gears will benefit from the same construction even if to a lesserdegree. FIGS. 42-44 show pinions with no inner cylinder. This may insome configurations provide enough bending stiffness and torsionalflexibility to provide adequate load sharing between all pinions.

Backlash is detrimental to the performance and precision of gearboxreducers in many applications such as robotics.

Disclosed here are mechanisms and constructions for equalizing the loadsharing between four or more pinions in a differential planetarygeartrain. Also disclosed here are mechanisms and constructions forreducing or eliminating backlash in a differential geartrain.

One object of the present device is to simplify and reduce the cost ofthe assembly through the elimination of the need for a planet carrier toposition the gears axially in the assembly. This is accomplished inembodiments of the device, through the use of opposite helical gears oneither end of the planet gears. Straight gears may also be used, asdisclosed below in relation to FIGS. 54-56. These gears are alsodesigned with a taper which has the benefit of making them easier topull out of a mold. The helical gears are beneficial for smoothoperation while the taper allows easy removal from an injection mold forlow cost production.

The symmetric construction of the design together with the opposinghelical gears on either side of the center plane and/or the tapering ofthe planets on either side of the center plane eliminates the need for aplanet carrier to keep the planets axially positioned. The symmetricalconfiguration of embodiments of the device eliminates the need for aplanet carrier because twisting of the planets perpendicular to theiraxis is virtually eliminated. This allows the planets to be hollow whichreduces weight and allows them to compress radially under radialpreloading to remove backlash while preventing binding during heatexpansion as shown in FIG. 51.

FIG. 47 shows axial forces on a planet gear of FIG. 38 in a gearbox ofFIG. 27, when the teeth of the planet gear are rotating in the directionshown by curved arrow 178. The directions of axial forces are shown byarrows 180. FIG. 48 is an isometric cutaway view showing the load pathson the planet 123 leading to the axial forces shown in FIG. 47. Thecircumferential forces of the load paths are shown by arrows 182. Allbut one planet and the retaining rings and bearings are omitted in FIG.48.

Load sharing can be provided for example by outer gear deformation, asshown schematically in FIG. 49, by sun ring deformation, as shownschematically in FIG. 50, or by planet gear deformation, as shownschematically in FIG. 51.

While the embodiment shown above has a sun input and outer output andreference ring gears, an outer input ring and sun output and referencegears is another possible arrangement.

Some properties of selected materials are shown in the below table.

TABLE 2 Yield Tensile Strength/ Strength Modulus Stiffness/ (ksi) (ksi)Density Strength/ Density Yield Tensile (g/cm3) Stiffness Strength/Material Strength Modulus Density Strength/ Stiffness/ Material (ksi)(ksi) (g/cm3) Stiffness Density Delrin 11 450 1.42 0.0244 0.0172 Nylon10.5 365 1.14 0.0288 0.0252 Glass- 28.3 1490 1.38 0.0190 0.0138Reinforced Nylon PEEK 16 540 1.3 0.0296 0.0228 Carbon Filled 40.6 32001.4 0.0127 0.0091 PEEK Glass Filled 29 1700 1.52 0.0171 0.0112 PEEK 606140 10000 2.7 0.0040 0.0015 Aluminum 4140 Steel 70.3 29700 7.85 0.00240.0003 Maraging 245 27600 8 0.0089 0.0011 Steel

FIGS. 52 and 53 are bar charts showing some material properties fromtable 2 in visual form.

FIG. 52 shows material strength to stiffness to density ratio. FIG. 52shows clear difference between plastics and metals when comparing thisratio. For this application it is better to have higher strength withlower stiffness, and lower density. FIG. 53 shows a material strength tostiffness ratio.

In an embodiment, pinions may preferably have a ratio of yield strengthto stiffness of greater than 0.010.

In an embodiment, pinions may have a ratio of torsion twist stiffness tobending stiffness of less than 1.

FIGS. 54-56 show another example of a differential gearbox. Unlike theembodiment in FIG. 27, this one has straight cut gears.

The gearbox comprises one or more sun gears 114 in geared contact withplural planet gears 123, the plural planet gears each in contact withdifferent ring gears 110 and 111. The planets may have a differentdiameter in contact with different ring gears. Thus, if the sun gear(s)114 is the input, a gear reduction may be obtained by the planetsdifferentially moving the ring gears. One or more ring gears may befixed and one or more may be an output. There may also be one or moreadditional floating sun gears contacting parts of the planets that arenot contacted by the sun gear(s) which are providing input torque.

As with other examples shown in this document, in different embodimentsall these input, output, fixed and optional floating gears may bearranged differently. For example, there may be output and fixed sungears and input ring gear(s) and optional floating ring gear(s). Theinput and output gears could be switched in any embodiment to changefrom a reducer to an increaser.

The gears may be arranged in an axially symmetric arrangement with axialcentering provided as described below, avoiding the need for a planetarycarrier.

The gears in the particular embodiment shown in FIGS. 54-59 use straightcut (spur) gears for both the inner planetary gear teeth 101 and theouter planetary gear teeth 102, and all the other gear teeth that meshwith these gear teeth. Helical gears as disclosed in other embodimentsin the incorporated material may also be used.

FIG. 59 shows a conical taper plus a profile shift, and a gear tooth 508with a profile shift is more specifically shown in FIG. 82. The taperuses both a conical taper of the gear as well as a taper of the geartooth involute profile by means of a profile shift that is varied acrossthe width of a tooth. This allows the gear tooth contact to behave in asimilar way to a conical taper, but retain the involute profile'sability to have proper gear tooth contact and conjugate motion. Theprofile shift acts to offset the involute profile in a positivedirection on one axial direction of the tooth and in a negativedirection on the axial direction. The meshing gear has the same shiftexcept in opposite directions. Profile shift is often used in gears butwill be constant across a tooth (partly due to manufacturing concerns)This can help make up for small errors in tooth contact or can help tooptimize a gear set. The magnitude of this is typically quite small, butcan be of similar magnitude as we are using. The profile shift acrossthe tooth width is nominally linear, but could possibly be non-linearacross the tooth width, providing the positive shift on one toothmatches the negative shift on the matching tooth. This profile shiftchanges across the depth of the tooth can be described as a tapering ofthe tooth itself along the axial direction which is independent of theconical taper of the gears. The profile shift may be applied to any ofthe gears. A suitable profile shift is shown for example inPCT/IB2018/055087, the content of which is hereby incorporated byreference.

One or both taper effects may be combined with shims between any one ormore of symmetrical output ring halves, output rings and fixed rings orsymmetrical sun gear halves to allow preload adjustment of the bearingsand of the gears to reduce or eliminate backlash.

The double helical or herringbone gears of the embodiment of FIG. 27provide a stabilizing force to center the planets axially, withoutcausing a net axial load due to the symmetric design, and would alsoprovide such a benefit in the embodiment of FIG. 54. In the embodimentshown in this document, the planet gears and corresponding gears on thesun gear and ring gears are tapered. The taper also provides thisstabilizing force to center the planets, so helical gears are notrequired for axial centering.

As described above, helical gears with different angles (or helicalgears paired with spur gears), in combination with springs allowingchanges in an axial separation of the gears of each planet, may be usedto provide load sharing. This is particularly useful for allowing loadsharing between relatively inflexible gears. In this particularembodiment shown, spur gears are used only, and the planets are a singlepiece. Thus, the above load sharing mechanism is not operative in thisembodiment. The gears being spur gears and the planets being formed of asingle piece is not necessary to the operation of the gear reducer, butthese features simplify construction.

Load sharing may be provided in this embodiment by flexibility of theplanet gears. The planet gears may be hollow and flex radially (forexample changing in cross sectional shape from circular to slightlyelliptical) to achieve a reduction or elimination of backlash. This maybe facilitated by the use of a relatively flexible material to make thegears, and by the thinness of the walls of the hollow gears. The absenceof a planet carrier aids in flexibility. Otherwise shafts and bearingsassociated with a planet carrier device would impede radial flexing.

In a typical rigid gear system, some space between meshing gears forthermal expansion is required to prevent binding, but this space allowsbacklash. The flexibility of the planets in this embodiment means thatadditional space to allow backlash is not necessary. The planets may bepreloaded radially using shims between the elements holding the ringgears, which combined with the radial taper are adjusted during assemblyto reduce or eliminate backlash. The radial flexibility of the hollowgears allows thermal expansion to flex the gears, making them ovalshaped, for example as shown in FIG. 51, without causing binding.

The shims may be for example flat and may include a shim along the axialcentral plane of the gearbox, and others on either side of the bearingparts as shown in the figures below. The shims position the ring gearsaxially, which allows the tapered ring gears to engage with the taperedplanets, taking up backlash.

Shims may be applied to adjust relative axial position of any of theelements relative to the axial center plane or each other. To maintainsymmetry, it would generally not be desirable to change the axialposition of elements that straddle the center plane. Such elements couldhowever also be used as a reference to shim other elements, for examplein an embodiment (not shown) with floating sun gears on either axialside of the input sun gear, and connected to it with bearings, shimsanalogous to shims B and C described below could be used to adjust theaxial positions of the floating sun gears relative to the input sun gearand axial center plane.

The embodiment shown in FIGS. 54-56 has central output ring gearsseparated by a center shim 184, and fixed outer ring gears connected tothe output ring gears by bearings with inner shim 186 and outer shim188, described below. Center shim 184 adjusts the axial position of theoutput ring gears relative to each other and to the planets. Inner shims186 and outer shim 188 adjust the positions of the fixed ring gears 110relative to the output ring gears 111, and in combination with thecenter shim 184, relative to the planets. There is also shown in FIG. 54a sun gear shim 187. Sun gear shim 187 adjusts the axial position of twosun gear halves (sun gears) relative to each other. This can be used toadjust the backlash as well. Sun gear shim 187 is not shown in the otherfigures for this embodiment. Corresponding elements are symmetricallyarranged on the other side of the center plane, so only one side isdescribed. Different shim adjustments may be used on different sides,but typically this would not be done except to correct anotherasymmetry.

In this particular embodiment, the sun gear 114 is one piece andstraddles the axial central plane, and does not have a shim.Alternatively, the sun gear may be made of two pieces, similar to theoutput gear 111 of this embodiment, with a shim between the two pieces.

The embodiment shown may be formed by injection molding, for example outof plastic. Helical gears also provide a more consistent contact ratioand potentially less noise and transmission error, but may be moredifficult to injection mold. The gears shrink axially as well asradially after molding, which makes it harder to get accurate finalparts. A spur gear may therefore be easier to construct, and may make iteasier to get precision parts, which may compensate for the advantagesof helical gears.

As shown in FIGS. 57-59, the planets each include portions with twodifferent pitch diameters in order to differentially drive the fixed andoutput gears. To enable injection molding with no interference as theplanets leave the mold, for example if the mold is formed of two halvesseparated at an axial central plane of the planets, in the embodimentshown the portions with the two different pitch diameters have the samenumber of teeth, and the teeth of the portions are aligned, as shownmost clearly in the end view of a planet in FIG. 59.

The planets may also be formed by machining, or be machined afterinjection molding. The portions with the two different pitch diametershaving the same number of teeth, and the teeth being aligned, alsoallows easier machining.

The fixed and output ring gears, on the other hand, would have differentnumbers of teeth in this embodiment to be driven differentially by theplanet gears.

Metal gears, for example of steel, may also be used. To compensate forthe greater stiffness, thinner walls may be used to allow radialflexibility.

As described below in relation to FIGS. 79-81, different planet gearsmay be arranged to be out of phase in terms of gear meshing. There maybe sets of gears, the gears of each set being at the same phase,different sets being at different phase, or all gears may be atdifferent phase. (Or they may all be at the same phase). For example, inan embodiment with 12 gears as shown here, there may be two sets of 6gears each, for example every second gear being in a set, or all twelvegears may be at a different phase of meshing at any given moment. Thisdistributes changes in forces due to the gear meshing or due toimperfections similarly aligned in different gears, so that they do notoccur all at once.

In the embodiment shown in FIGS. 54-56, dual bearings are includedbetween the fixed and output ring gears on each side of the axial centerplane. The bearings are each angular contact bearings and may bepreloaded using inner shim 186 and outer shim 188. The difference ofwidth between inner shim 186 and outer shim 188 may be used to preloadthe bearings and adding or subtracting from the width of both of innershim 186 and outer shim 188 equally may be used to determine the axialseparation between the fixed and output ring gears. This arrangement hasgood stability. The sets of bearings on each side of the axial centerplane may be preloaded separately. The set on the opposite side may bepreloaded using the shims symmetric to inner shim 186 and outer shim 188with respect to the axial center plane.

In an embodiment, the bearings may be made out of plastic. This allowsreduced weight of the bearings compared to metal bearings.

FIG. 54 shows an isometric cutaway view, and FIG. 55 shows an isometricview, of this embodiment. FIG. 56 shows an axial end view.

FIGS. 57-59 show an exemplary planet for the differential planetarygearbox if FIGS. 54-56. FIGS. 57 and 58 show isometric views and FIG. 59shows an axial end view.

Thus, there is provided a differential gearbox with a driving sun inputon the ID of the pinions and one or more fixed ring gears on the OD ofthe pinions and one or more output ring gears on the OD of the pinionswith a number of pinions greater than 5, 6, 7, 8, 9, 10, 11, 12, 13, 14,15, 16, 17, 18, 19, 20 pinions with pinions and/or inner or outer ringgears: having a yield strength-to-stiffness ratio of greater than 0.10;made of plain or fiber reinforced polymer resin; with pinions having atorsionally flexible section between inner and outer gears to provide atorsion twist stiffness-to-bending stiffness of less than 1. There maybe axial or helical slots in the cylindrical sections to increasetorsional flexibility. There may be symmetrical opposing tapered pinionsthat allow backlash adjustment by axial shimming or other fixedpositioning means for fixed and/or output ring gears. The gears may bestraight or helical. There may be two sets of opposed bearing racesintegrated into housing and output ring gear.

Example Planet Driven Actuator

Embodiments of an exemplary actuator as shown in FIGS. 60-67 include anintegrated symmetric differential gearbox coupled to a direct drivemotor with two stators. Rectangular boxes in FIG. 60 show the portionsof the gearbox corresponding axially to the output gear (in the axialcenter) and the fixed gears (above and below the output gear).Embodiments of the exemplary actuator further include the use of thepinions as permanent magnet carriers which serve the purpose of a rotorwithout the need for a separate rotating magnet carrier. Components maybe enclosed in a sealed housing 212 and rotate about a common axisindicated by reference numeral 290, and may be symmetric about the axialcenter plane indicated by reference numeral 292. One or moreelectromagnetic stators provide the input torque to the gearbox and maybe located outward from the axial center plane 292 of the device. Aircoils (that is, electromagnetic coils with no soft magnetic core) may beused in the stators, and are commutated to act on permanent magnetswhich are inserted into the pinion gears. The pinion gears may beradially preloaded by a floating rolling contact sun ring 294 shown inFIG. 61. The pinion gear teeth have two symmetric tooth profiles whichmay be one or more of helical, tapered, or have an involute profile. Thepinion teeth which are located at the axial center of the device,referred to as the inner pinion teeth 201, mesh with the output ringgear teeth 274 of output gear 211 at the axial center plane of thedevice. The pinion teeth located outward from the axial center of thedevice, referred to as the outer pinion teeth 202, mesh with the teeth276 of two fixed input ring gears 210 which are located axially outwardfrom the axial center of the device. The fixed ring gears 210 aresecured together, such as through the axial center of the device. Adifferent gear tooth ratio between the pinion and the fixed ring gear,and pinion and the output ring gear, is what causes the output ring tospin at a reduced speed relative to the orbiting speed of the pinions.The non-limiting exemplary embodied shown here has a 10:1 ratio and canbe scaled accordingly to provide multiple ratios for differentapplications. FIG. 62 shows the device without the motor inputcomponents.

A non-limiting exemplary embodiment shown here comprises 8 planetarygears that have no sun gear input. A high number of planetary piniongears 223 allows for a large torque capacity from the device. Loadsharing is accomplished by applying magnetic force directly to thepinions 223 together with the rolling contact free spinning sun ring 294which preloads the pinons against the fixed rings 210 and output rings211 while allowing each pinion to find it is ideal circumferentialposition so all of the pinions are reasonably equally loaded when torqueis applied through the electromagnetic stators. The pinons are kept ingear engagement with the outer rings 210 and 211 by using a rollingcontact sun ring 294 closest to the radial center of the device. It isworth noting that the two fixed ring gears 210 are attached to eachother in this embodiment as if they are one piece by means of acylindrical member 296 that connects the fixed ring gears through thelarge center through hole of the actuator. The use of a tractioninterface between the rolling contact sun ring 294 and pinions 223allows the pinions 223 to each find their own individual position withrespect to the fixed and output gear rings. This may allow each of thepinons to more equally share the load that results from torque transferof the device than if the inner ring was geared, for example.

The rotation of the planetary gears 223 is accomplished by the influenceof an electromagnetic force produced by electromagnetic conductor coils298 acting on permanent magnets 300 which are secured in the piniongears 223 through both axial ends of the gears by an attracting magneticforce to a ferrous cylinder 302, such as but not limited to steel oriron, centered in the planets. In a non-limiting embodiment, motorstators use air coils 298 in combination with back iron 304 and a smallenough air gap between magnets and back iron that a reasonably hightorque can be achieved while at the same time eliminating passivecogging effects because there are no steel posts in the stator. Steelposts may be used as well for different torque effects. Alternatively,the electromagnets may have soft magnetic posts (not shown). If aircoils are used, or if electromagnets with unusually small soft magneticposts are used, the stator may use a soft magnetic material backiron toimprove the efficiency of the electromagnetic coils with minimal or nocogging.

The laminated back iron 304 in the motor stators may be made up ofalternating layers of laminates and protruding aluminum fins 306 whichextend through openings in the fixed input ring gear 210/housing.Manufacturing is simplified because the back iron 304 is made of spiralsof steel (represented by concentric rings in FIG. 63 which shows anupper view of components of the upper motor) and does not require anyalignment of pole sections. Periodic layers (such as every second layer)of the concentric winding can be steel and serve as a flux path betweenair coils, but is preferably hard anodized aluminum or some other highheat conductivity material such but not limited to aluminum.

In this non-limiting configuration the aluminum provides a heat pathdirectly from the coil to the heat extraction surfaces on the outside ofthe actuator. Heat is transferred from the copper conductors of the aircoils 298 into the back iron 304 and dispersed at openings in thehousing as air passes through the fins 306. An air fence component 308with extrusions 312 formed in a serpentine pattern overlaps withopenings 310 in the housing to guide air from an input past the fins 306and out of the device, best shown in FIG. 65. Arrows 316 show air flowthrough cavities of the air fence 308.

FIG. 64 shows a bottom view of components of the upper motor. FIG. 64shows a circumferential offset of cavities defined by the air fence 308above aluminum fins 306.

In order to simplify manufacturing, planet gears can be made to beinjection moldable such as from plastic. A configuration of the planetgears can include continuous features such as the helical teeth alongtheir axes without an undercut as shown in FIG. 66 to ease injectionmolding.

FIG. 67 shows an exploded view of an upper half of the gearbox of FIG.60.

The taper required for injection molding then requires that the maximumaddendum diameter of the axially outer section of a planetary gear mustbe smaller than the minimum addendum diameter of the axially innersection. A preferred embodiment of the planet gears would include anumber of teeth in the axially outer section to be either a factor of orequal to the number of teeth of the axially inner gears. In anembodiment with injection moldable planet gears, a geared floating innerring can be used to preload pinion gears and keep them in engagementwith outer rings. A geared floating inner ring may have the advantage oftransferring torque from pinions that are under high magnetic force topinions that are in between phases of EM force input. It is worth notingthat if the number of teeth on the axially outer section is a factor ofthe number of teeth on the axially inner section the whole gear canstill be injection molded with step change at the start of the tooth onthe axially inner one that does not align with a tooth on the axiallyouter.

An embodiment includes seals that are added along the inputs and outputsof the motor to protect inner components from dust and particulate.

The tapered gears allow geared parts to be removed from an injectionmold or other process such as compacted powder process such as powderedmetal. The taper may have a different helix angle on leading andtraining faces of each gear tooth and/or taper of addendum and dedendumso that the gear releases from the mold with no binding. Othercharacteristics of the tapered tooth design allow the involute shape tooperate correctly.

There may be a rolling feature (e.g., cylindrical section) between theinner and outer gears on the pinion 223. This provides a mid-forceposition for the preload provided by the rolling sun ring 294.

This also allows the steel pinion core 302 width to be wide enough(axially long enough) so rolling contact of sun ring 294 on plasticpinion race (rolling feature) transfers compression load to steel core302 rather than to magnets 300 on either end of steel core. Steel hasmuch better strength and endurance than typical PM magnet material sothis will contribute to increased service life.

The magnets 300 are preferably not stressed by the preload of the sunring 294, so the PM magnets can be magnetically attracted to andtherefore attached to the axial ends of the steel cores withoutcontacting the ID of the ends of the plastic pinion gears. A cylindricalboss on the inner axial end of the PM's will fit into a circular bore inthe end of the steel core to center the PM's.

The PM's 300 may have a larger OD beyond the axial ends of the pinions223 to provide a larger magnetic pole to interact with theelectromagnetic stators.

Assembly sequence may be as follows and is enabled by a symmetric splitoutput ring 210.

If the pinons 223 are placed into the ring gears as a prior step, thesun ring contacts would interfere with the pinions preventing assemblyof the sun ring. Likewise, if the sun ring and pinions are assembledfirst, the output ring would interfere with the pinons and would preventassembly. By using a split output ring gear 210, however, it is possibleto assemble the pinions 223 and sun ring 294 or sun ring assembly first,and then to assemble each of the two output ring gear halves 211together toward the center plane. The two tapered fixed gear rings 210are then assembled. This prevents any interference during assembly andallows a zero backlash final result.

Note that to achieve a true zero backlash final assembly it is providedthat the axial position of the fixed rings 210 can be adjusted relativeto the output ring gear 211 assembly. The problem with that solution isit makes an integrated bearing race 314 between the fixed and outputrings very difficult to implement. Instead, an embodiment uses anexpandable sun ring 318, as shown below, which can provide gear preloadat a range of radial positions for the pinion center axes. FIG. 68 showsa planet 223 having a shape suitable to be used with an expandable sunring 318 for example as shown in FIG. 69. The planet 232 may have amagnet 300 arranged with clearances 320 between the magnet and theplanet gear teeth portion, and between the magnet and radially separatedportions of the steel core 302. A line 322 indicates a tooth angle ofthe inner 201 and outer 202 teeth.

A replaceable/adjustable shim (not shown) is provided in the assemblybetween the two output ring gear halves to adjust their relative axialposition to each other at the center plane. This shim can be areplaceable ring or ring sections that can be inserted during assembly,or it can be a rotating ring with ramps or threads that can be adjustedvia rotation around the actuator center axis after assembly of all thecomponents but before the two output ring halves are secured togethersuch as by tightening together with bolts. By creating the assembly sothe output ring gears 210 have backlash when at the maximum shimadjustment thickness, the fixed 210 and output 211 ring gears andpinions 223 and expandable sun gear 318 can be loosely arranged butassembled close to their final positions. At this point, the preload ofthe sun ring 318 will push the outer pinion gears 202 and fixed ringgear 210 teeth into engagement with no backlash. Reducing the axialdistance between the output gear rings 211 will then remove the backlashfrom the inner gears 201 and output rings 211. To allow a single shimstack (between only the two halves of the output ring gears) to adjustthe backlash in the fixed and output gear meshes, requires that thepinons can be moved outward as a result of the expanding sun ring 318,and that the taper on the inner gears 201 of the pinion 223 (and theoutput gear 211) be of a greater taper angle. This way the axialadjustment of the output ring gears 211 has more effect on the innergear mesh between inner teeth 201 and output gears 211 than the outergear mesh between outer teeth 202 and fixed gears 210. As a result, therelative position of the fixed ring gear 210 on an end and the outputring gear 211 half on that end, do not need to be changed. This allowsthe bearing races 314 to be molded or machined into the fixed and outputgears because there is no change in this relative position requiredduring backlash adjustment. The only thing that must be considered inthis adjustment strategy is that the relative position of the two fixedring gears 210 to each other will need to be shimmed/adjusted at thesame time as the two output ring gears 211 are adjusted. This adjustmentcould, however, be a result of the compliance of the housing 212 whichsupports and secures, including via connecting portion 296, the twofixed rings 210 together.

Adjustable sun ring pressure via axially spring two-piece sun ring witha mechanical spring (or repelling magnet ring) pushing axially outwardwill compensate for thermal expansion and gear surface wear whilereducing or eliminating backlash. This allows for a greater tolerancerange on the diameters of the sun ring and pinion contact races, as wellas the tolerance of the gear faces.

An axially expandable sun ring 318, as shown in FIG. 69, provides thecapability to adjust the preload force of the sun ring races on thepinion races. This can be done by a number of means including, forexample, a magnetic coil 324 on both axial ends of the axiallyexpandable sun ring assembly as shown below. The electromagnetic coilsin this example are fixed to the actuator housing. By powering them to alevel that is proportional to the output torque of the actuator, thepreload of the pinions and ring gears can be adjusted at all times toensure a zero-backlash gear interface and the lowest possible friction.

In an embodiment, shown in FIG. 70, a power-off safety brake isintegrated into the gearbox as follows. A separation spring 326 isstrong enough to create a high level of friction between the sun ring318 and pinon contact face 328. This friction also results in higherfriction between the gear mesh surfaces making the gearbox verydifficult or impossible to backdrive. This results in a power-off brakethat prevents rotation of the actuator when power is lost. To releasethe brake, an actuation means, such as an electromagnetic coil 324exerts a force in the opposite direction. The level of this force canthen be adjusted to relieve enough of the spring force to eliminateunnecessary friction, but not so much as to cause backlash at any giventorque output. For clarity, the greater the EM force on the sun ringassembly, the lower the preloading force on the pinions. When power islost to the actuator, the EM coils become unpowered and the mechanicalspring provides the preload force which is then high enough to make thegearbox less or completely non-backdrivable.

Features of embodiments of this planet driven actuator may include: thefixed ring gears are secured together, such as through the axial centerof the device; planet gears would include a number of teeth in theaxially outer section to be either a factor of or equal to the number ofteeth of the axially inner gears; serpentine air fence causes air to pascircumferentially between cooling fins as it flows from the center axisoutward; a non-geared sun ring preloads pinons but allows them to floatfor equal torque transfer; the sun ring can be geared; the sun ring mayalso distributes torque between pinions when some pinons are no beingpulled along by stator; and the sun ring can be used as a brake member(not shown) by applying braking force to sun ring (whether geared ornot) with a brake device (not shown). An alternative brake using a sunring is shown in FIG. 70.

Planetary Bearing

A rotor can be coupled to a gearbox and make use of the planet gears asa bearing with the sun as the interface. Six non-limiting configurationsof gearboxes (71-78) are presented to demonstrate the configuration ofgears such that the planets can act as bearings. This can work for anynumber of planets greater than 3. FIGS. 71-78 cover configurations ofsymmetric planetary gearboxes with either axially inner or axially outersun interfaces of many possible gear ratios. This principle works in anyplanetary system with an input gear and a stationary gear on opposingradial sides (ex. left and right of planet gear in FIGS. 71-78) of aplanetary gear.

Each of FIGS. 71-78 shows a 2D section view of a respective embodiment.Gear teeth are not shown for simplification. Each shows a planet 400interfacing with an input gear 402, a stationary gear 404, and an outputgear 406. A dotted line 408 indicates a central axis about which theinput and output gears rotate.

FIG. 71 shows an embodiment using a sun input 402 interfacing withaxially inner planet gear teeth and outer stationary gear 404interfacing with axially outer planet gear teeth.

FIG. 72 shows an input ring gear 402 interfacing with the axially inner(or outer) planet gear teeth and a stationary sun gear 404 interfacingwith the axially outer planet gear teeth (and curving outward to looparound the input ring gear), and an output sun gear 406 interfacing theaxially inner planet gear teeth. FIG. 72 is the same as FIG. 71 but withwhat is radially inner and what is radially outer inverted.

FIG. 73 shows a sun input 402 interfacing with axially outer planet gearteeth and outer stationary 404 interfacing with axially outer planetgear teeth.

FIG. 74 is the same as FIG. 73 but with what is radially inner and whatis radially outer inverted.

FIGS. 75 and 76 have a stationary portion surrounded by a moving part.This would not work for continuous motion, but would be perfectlyallowable for a limited range or motion joint, i.e., the moving partwould have a slot part of the way around the circumference which allowsthe stationary part to be grounded. These would also be a speedincreaser if the axially inner planet gear is larger than the axiallyouter planet gear. This could also be reversed to change it to areducer, although this may make it more difficult tomanufacture/assemble the planet gear.

FIG. 75 shows an outer input gear 402 interfacing with axially innerplanet gear teeth and stationary sun gear 404 interfacing with axiallyinner planet gear teeth.

FIG. 76 shows an outer input gear 402 interfacing with axially innerplanet gear teeth and stationary sun gear 404 interfacing with axiallyouter planet gear teeth.

FIG. 77 shows an outer input gear 402 interfacing with axially outerplanet gear teeth and a stationary sun gear 404 interfacing with axiallyinner planet gear teeth.

FIG. 78 shows an outer input gear 402 interfacing with axially outerplanet gear teeth and a stationary sun gear 404 interfacing with axiallyouter planet gear teeth.

Out of Phase Gears

One potential drawback when using the differential gearbox is theadditional gear meshes when compared to a three or four planet system.When the planets are all in phase, gears make contact with the nexttooth at the same time, and with additional planets, the potential fornoise and vibration are increased. By keeping the planets out of phase,the additional initial tooth contacts can be distributed more evenly,significantly reducing the ‘cogging feel’ of an in-phase system. In oneexemplary embodiment, there are 14 planets and thus 14 initial contactsfor each tooth. In a single phase system, all of these contacts occur atone point in time, while in a 7 phase system, only two contacts occur ata single point in time, but the contacts occur 7× as often.

In one embodiment, shown in FIG. 79, and FIG. 80, there are 14 planets123 positioned evenly around the sun 114. Each planet is identical andthere are 7 unique phases of planet. Opposite planets are in the samephase. The gear diameters and tooth numbers are provided in Table 3.

TABLE 3 Scaled Diameter Teeth Sun 89.18 114 Outer 129.86 166 Ring Stator126.47 194 Ring Planet 20.34 26 Large Planet 16.95 26 Small Sun 2 92.57142

This can be separated into two traditional planetary gearsets: an innergearset 500 and outer gearset 502, each with a sun gear, planet gear,and ring gear. Each gearset has a constant pitch or module, which may bedifferent from the other gearset. Note that the differential gearsetdoes not make use of a sun gear on the outer gearset. The virtual sungear would fit into the system, but is not required due to the balancednature of the system.

TABLE 4 Inner Gearset Outer Gearset Diameter Teeth Diameter Teeth Sun89.18 114 92.57 142 Ring 129.86 166 126.47 194 Planet 20.34 26 16.95 26

The two gearsets are designed such that diameters and tooth numbers inthe system are compatible in:

-   -   The number of planets    -   The radial and tangential position of the planets    -   The phase of the planets

Individually, the gearsets follow known rules for planetary geartrains,with the sum of the number of teeth in the sun and ring gears beingevenly divisible by the number of planets. This ensures that the planetsare evenly spaced around the sun.

$ {{Inner}\mspace{14mu} {Gearset}}\;arrow\frac{114 + 166}{14}  = 20$$ {{Outer}\mspace{14mu} {Gearset}}\;arrow\frac{142 + 194}{14}  = 24$

The radial position of the planets is controlled by the relative scalesof the two gearsets, ensuring that the planet gear axes fall on the sameradius.

The phase of the planets is kept compatible by ensuring that the twogearsets have the same number of unique phases. The number of uniquephases in the system is calculated by dividing the number of planets bythe greatest common factor of the number of teeth on the sun gear, ringgear, and the number of planets.

$ {{Inner}\mspace{14mu} {Gearset}}\;arrow\frac{14}{{GCF}( {114,166,14} )}  = {\frac{14}{2} = 7}$$ {{Outer}\mspace{14mu} {Gearset}}\;arrow\frac{14}{{GCF}( {142,194,14} )}  = {\frac{14}{2} = 7}$

The phases are organized such that similar planet phases are distributedevenly around the sun. In the example above, there are 2 planets foreach unique phase. The two planets in one phase will be located 180°from each other.

With the same number and position of unique gear phases, the system canoperate. If only the position were compatible, the system could operateonly if there were planet configurations specified for each individualgear position. Each planet configuration would use identical gears, butthe rotational alignment between the inner and outer gears would beunique for each phase pair.

In an embodiment, the number of teeth on the planet gears may either beequal, or the number of teeth on one planet may be an integer multipleof the number of teeth of the other planet. This provides a number ofadvantages.

First, the planet may be more easily manufactured as a single pieceusing methods including, but not limited to injection molding. It isbeneficial to have the planets as a single piece in order to betterallow for load sharing mechanisms in the system. One of the load sharingmechanisms that the exemplary system relies on is the radial compressionof the planets to ensure gear tooth contact on both the sun and theouter ring. One of the ways to ensure this occurs is to include a largethrough hole in the planet, allowing for some additional radialflexibility. If the planet were to be constructed of multiple pieces,the method of fastening them together would require significantly morematerial and thus result in significantly more stiffness.

Second, the system may be assembled with the planets as a single piece.In any planetary gearset, one or more of the gears must be insertedaxially. In one process, planets could be placed around the sunradially, and the outer ring may be inserted axially around the planets.For the differential gearbox, two planetary gearsets are joined axiallyand if the planet is a single piece, there are limitations to theassembly method. By constraining the number of teeth on the planetgears, there are significantly more configurations that may beassembled. Without this constraint, some configurations cannot beassembled, as the outer planet gear teeth would prevent the axialinsertion of a gear into the inner gearset. This is applicable for spurand helical gear teeth. See the additional specification filed herewithfor images of the different configurations (sun input, ring input,etc.).

For most reasonable gear ratios, the two planet diameters must besimilar. If the addendum of the smaller planet gear is larger than thededendum of the larger gear, an undercut exists, meaning that a gearcould not be axially inserted into position onto the inner gearset.

In FIG. 81, it is clear that the two planets shown are out of phase. Therespective lines 504 running between the gear center axes and the centeraxis of the device intersects the gear teeth 506 at a different point onthe nearest tooth, falling near the center of the sun tooth on the leftgear, and falling to one side of the sun tooth on the right gear.

By using the principles described above, there are a number of uniquesolutions that may be used in a differential gearbox. A list of thesesolutions is shown below, but the scope of the claims is not limited tothese specific solutions. Additional solutions exist with differentnumbers of gear teeth and each solution may be geometrically scaled tosuit any diameter, keeping the number of teeth constant. Specificconfigurations can be determined by solution of known equations appliedfollowing the principles in this disclosure.

TABLE 5 Number Planet Planet Outer Stator Planet Planet Gear of SunOuter Ring Stator Ring Large Small Sun 2 Sun Ring Ring Large Small Sun 2Ratio Planets Diameter Diameter Diameter Diameter Diameter DiameterTeeth Teeth Teeth Teeth Teeth Teeth 7.220339 26 137.6667 165.6667161.6667 14 10 141.6667 118 142 194 12 12 170 3.61017 26 196.6667236.6667 226.6667 20 10 206.6667 118 142 272 12 12 248 2.707627 26255.6667 307.6667 291.6667 26 10 271.6667 118 142 350 12 12 326 13.2372926 118 142 140 12 10 120 118 142 168 12 12 144 4.211864 26 177 213 20518 10 185 118 142 246 12 12 222 2.922518 26 236 284 270 24 10 250 118142 324 12 12 300 2.54049 26 275.3333 331.3333 313.3333 28 10 293.3333118 142 376 12 12 352 3.20904 26 216.3333 260.3333 248.3333 22 10228.3333 118 142 298 12 12 274 5.214689 26 157.3333 189.3333 183.3333 1610 163.3333 118 142 220 12 12 196 5.253687 25 150.6667 182.6667 176.666716 10 156.6667 113 137 212 12 12 188 3.233038 25 207.1667 251.1667239.1667 22 10 219.1667 113 137 287 12 12 263 2.559489 25 263.6667319.6667 301.6667 28 10 281.6667 113 137 362 12 12 338 13.33628 25 113137 135 12 10 115 113 137 162 12 12 138 4.243363 25 169.5 205.5 197.5 1810 177.5 113 137 237 12 12 213 2.944374 25 226 274 260 24 10 240 113 137312 12 12 288 2.727876 25 244.8333 296.8333 280.8333 26 10 260.8333 113137 337 12 12 313 3.637168 25 188.3333 228.3333 218.3333 20 10 198.3333113 137 262 12 12 238 7.274336 25 131.8333 159.8333 155.8333 14 10135.8333 113 137 187 12 12 163 4.569328 24 173.9231 211.9231 203.9231 1911 181.9231 119 145 241 13 13 215 2.893908 24 247.1539 301.1539 285.153927 11 263.1539 119 145 337 13 13 311 2.70775 24 265.4615 323.4615305.4615 29 11 283.4615 119 145 361 13 13 335 3.89916 24 192.2308234.2308 224.2308 21 11 202.2308 119 145 265 13 13 239 14.62185 24 119145 143 13 11 121 119 145 169 13 13 143 3.452381 24 210.5385 256.5385244.5385 23 11 222.5385 119 145 289 13 13 263 7.920168 24 137.3077167.3077 163.3077 15 11 141.3077 119 145 193 13 13 167 3.133253 24228.8462 278.8462 264.8462 25 11 242.8462 119 145 313 13 13 287 5.68627524 155.6154 189.6154 183.6154 17 11 161.6154 119 145 217 13 13 19113.44444 24 108 132 130 12 10 110 108 132 156 12 12 132 7.333333 24 126154 150 14 10 130 108 132 180 12 12 156 5.296296 24 144 176 170 16 10150 108 132 204 12 12 180 4.277778 24 162 198 190 18 10 170 108 132 22812 12 204 3.666667 24 180 220 210 20 10 190 108 132 252 12 12 2283.259259 24 198 242 230 22 10 210 108 132 276 12 12 252 2.968254 24 216264 250 24 10 230 108 132 300 12 12 276 2.75 24 234 286 270 26 10 250108 132 324 12 12 300 2.580247 24 252 308 290 28 10 270 108 132 348 1212 324 13.44444 26 108 132 130 12 10 110 117 143 169 13 13 143 7.33333326 126 154 150 14 10 130 117 143 195 13 13 169 5.296296 26 144 176 17016 10 150 117 143 221 13 13 195 4.277778 26 162 198 190 18 10 170 117143 247 13 13 221 3.666667 26 180 220 210 20 10 190 117 143 273 13 13247 3.259259 26 198 242 230 22 10 210 117 143 299 13 13 273 2.968254 26216 264 250 24 10 230 117 143 325 13 13 299 2.75 26 234 286 270 26 10250 117 143 351 13 13 325 2.580247 26 252 308 290 28 10 270 117 143 37713 13 351 13.55357 25 103.3846 127.3846 125.3846 12 10 105.3846 112 138163 13 13 137 2.601191 25 241.2308 297.2308 279.2308 28 10 259.2308 112138 363 13 13 337 2.772321 25 224 276 260 26 10 240 112 138 338 13 13312 2.992347 25 206.7692 254.7692 240.7692 24 10 220.7692 112 138 313 1313 287 3.285714 25 189.5385 233.5385 221.5385 22 10 201.5385 112 138 28813 13 262 3.696429 25 172.3077 212.3077 202.3077 20 10 182.3077 112 138263 13 13 237 4.3125 25 155.0769 191.0769 183.0769 18 10 163.0769 112138 238 13 13 212 5.339286 25 137.8462 169.8462 163.8462 16 10 143.8462112 138 213 13 13 187 7.392857 25 120.6154 148.6154 144.6154 14 10124.6154 112 138 188 13 13 162 7.398058 23 120.1667 148.1667 144.1667 1410 124.1667 103 127 173 12 12 149 3.699029 23 171.6667 211.6667 201.666720 10 181.6667 103 127 242 12 12 218 2.774272 23 223.1667 275.1667259.1667 26 10 239.1667 103 127 311 12 12 287 13.56311 23 103 127 125 1210 105 103 127 150 12 12 126 4.315534 23 154.5 190.5 182.5 18 10 162.5103 127 219 12 12 195 2.994452 23 206 254 240 24 10 220 103 127 288 1212 264 2.603021 23 240.3333 296.3333 278.3333 28 10 258.3333 103 127 33412 12 310 3.288026 23 188.8333 232.8333 220.8333 22 10 200.8333 103 127265 12 12 241 5.343042 23 137.3333 169.3333 163.3333 16 10 143.3333 103127 196 12 12 172 3.181598 24 210.7143 260.7143 246.7143 25 11 224.7143118 146 314 14 14 286 4.639831 24 160.1429 198.1429 190.1429 19 11168.1429 118 146 242 14 14 214 14.84746 24 109.5714 135.5714 133.5714 1311 111.5714 118 146 170 14 14 142 2.938559 24 227.5714 281.5714 265.571427 11 243.5714 118 146 338 14 14 310 3.959322 24 177 219 209 21 11 187118 146 266 14 14 238 2.749529 24 244.4286 302.4286 284.4286 29 11262.4286 118 146 362 14 14 334 8.042373 24 126.4286 156.4286 152.4286 1511 130.4286 118 146 194 14 14 166 3.50565 24 193.8571 239.8571 227.857123 11 205.8571 118 146 290 14 14 262 5.774011 24 143.2857 177.2857171.2857 17 11 149.2857 118 146 218 14 14 190 2.757202 22 240.9231298.9231 280.9231 29 11 258.9231 108 134 332 13 13 306 3.190476 22207.6923 257.6923 243.6923 25 11 221.6923 108 134 288 13 13 262

These parameters for the gearbox, using two sizes of radially outerplanets which alternate as you look around the circle (A,B,A,B,A . . .), while each of the radially inner planets are of the same size, arebelieved to work: (P=planet, R=ring, in =inner, out=outer).

TABLE 6 Diameter Teeth Total Soln # Ps Sun Outer R P Out P Out 2 P InSun Out R P Out P Out 2 P In Pitch Ratio Error 1 14 105 168 21.9 18.618.6 350 560 73 62 62 3.33 2.67 0.102 2 10 98 161 29.4 23.1 17.5 140 23042 33 25 1.43 2.56 0.144 3 14 114.8 173.6 12.2 10.4 28 574 868 61 52 1405 2.95 0.156 4 10 79.9 159.8 27.5 22.3 23.78 215 430 74 60 64 2.69 20.193 5 10 86 172 29.6 24 25.6 215 430 74 60 64 2.5 2 0.198 6 16 103.2163.2 19.8 18.3 16.5 344 544 66 61 55 3.33 2.72 0.202 7 16 103.2 163.219.8 18.3 16.5 344 544 66 61 55 3.33 2.72 0.202 8 10 81.35 162.7 28 22.724.22 215 430 74 60 64 2.64 2 0.203 9 16 116.8 156.8 15.8 14.4 13.8 584784 79 72 69 5 3.92 0.227 10 16 116.8 156.8 15.8 14.4 13.8 584 784 79 7269 5 3.92 0.227 11 12 105.6 172.8 15.4 13.6 30 528 864 77 68 150 5 2.570.253 12 18 126.9 189 20.4 18.9 17.4 423 630 68 63 58 3.33 3.04 0.255 1318 126.9 189 20.4 18.9 17.4 423 630 68 63 58 3.33 3.04 0.255 14 18 126.9172.8 14.1 12.3 17.4 423 576 47 41 58 3.33 3.76 0.255 15 18 126.9 172.814.1 12.3 17.4 423 576 47 41 58 3.33 3.76 0.255 16 18 132.75 180 14 12.518.5 531 720 56 50 74 4 3.81 0.265 17 18 132.75 180 14 12.5 18.5 531 72056 50 74 4 3.81 0.265 18 24 150.6 194.4 16.5 9.9 15 502 648 55 33 503.33 4.44 0.267 19 24 150.6 194.4 16.5 9.9 15 502 648 55 33 50 3.33 4.440.267 20 14 119 173.6 15.8 11.6 24.2 595 868 79 58 121 5 3.18 0.273 2110 104 196 24.4 20.8 36.4 260 490 61 52 91 2.5 2.13 0.273 22 22 116.8158.4 13.2 9.4 14.4 584 792 66 47 72 5 3.81 0.281 23 12 104.4 183.6 21.320.1 28.2 348 612 71 67 94 3.33 2.32 0.31 24 26 121.2 156 11.4 9.9 10.8404 520 38 33 36 3.33 4.48 0.324 25 26 121.2 156 11.4 9.9 10.8 404 52038 33 36 3.33 4.48 0.324 26 20 146.02 189 18.9 12.5 14.02 479 620 62 4146 3.28 4.4 0.327 27 20 146.02 189 18.9 12.5 14.02 479 620 62 41 46 3.284.4 0.327 28 20 143.7 186 18.6 12.3 13.8 479 620 62 41 46 3.33 4.4 0.33229 20 143.7 186 18.6 12.3 13.8 479 620 62 41 46 3.33 4.4 0.332 30 20141.38 183 18.3 12.1 13.58 479 620 62 41 46 3.39 4.4 0.336 31 20 141.38183 18.3 12.1 13.58 479 620 62 41 46 3.39 4.4 0.336 32 12 100.56 161.524.3 17.2 19.44 269 432 65 46 52 2.67 2.65 0.342 33 24 118.2 158.4 13.512 11.4 394 528 45 40 38 3.33 3.94 0.343 34 24 118.2 158.4 13.5 12 11.4394 528 45 40 38 3.33 3.94 0.343 35 12 99.47 160 13.6 11.2 28.53 373 60051 42 107 3.75 2.64 0.344 36 12 111.9 180 15.3 12.6 32.1 373 600 51 42107 3.33 2.64 0.344 37 12 124.33 200 17 14 35.66 373 600 51 42 107 32.64 0.344 38 12 113.05 180.6 24.5 16.8 24.5 323 516 70 48 70 2.86 2.670.347 39 12 107.6 172.8 26 18.4 20.8 269 432 65 46 52 2.5 2.65 0.348 4010 120 195 21 15.5 37.5 240 390 42 31 75 2 2.6 0.353 41 10 96 156 16.812.4 30 240 390 42 31 75 2.5 2.6 0.353 42 20 127.08 169 20.8 11.7 11.7391 520 64 36 36 3.08 4.03 0.358 43 20 127.08 169 20.8 11.7 11.7 391 52064 36 36 3.08 4.03 0.358 44 20 146.63 195 24 13.5 13.5 391 520 64 36 362.67 4.03 0.358 45 20 146.63 195 24 13.5 13.5 391 520 64 36 36 2.67 4.030.358 46 20 136.85 182 22.4 12.6 12.6 391 520 64 36 36 2.86 4.03 0.35847 20 136.85 182 22.4 12.6 12.6 391 520 64 36 36 2.86 4.03 0.358 48 20117.3 156 19.2 10.8 10.8 391 520 64 36 36 3.33 4.03 0.358 49 20 117.3156 19.2 10.8 10.8 391 520 64 36 36 3.33 4.03 0.358 50 12 106.56 183.422.5 20.8 26.32 251 432 53 49 62 2.36 2.39 0.367 51 14 101.2 162.4 21.216.8 18.8 253 406 53 42 47 2.5 2.65 0.369 52 12 100.4 172.8 21.2 19.624.8 251 432 53 49 62 2.5 2.39 0.371 53 12 117 180 18.3 13.2 28.8 390600 61 44 96 3.33 2.86 0.379 54 8 87.96 158.9 22.7 18.8 32.63 248 448 6453 92 2.82 2.24 0.38 55 16 124.8 163.2 17.7 14.7 13.2 416 544 59 49 443.33 4.25 0.381 56 16 124.8 163.2 17.7 14.7 13.2 416 544 59 49 44 3.334.25 0.381 57 8 101.53 183.4 26.2 21.7 37.66 248 448 64 53 92 2.44 2.240.381 58 12 107.45 155.4 22.4 14.7 17.85 307 444 64 42 51 2.86 3.240.389 59 12 122.8 177.6 25.6 16.8 20.4 307 444 64 42 51 2.5 3.24 0.38960 12 138.15 199.8 28.8 18.9 22.95 307 444 64 42 51 2.22 3.24 0.389 6120 116.68 167 16.7 15.6 13.84 531 760 76 71 63 4.55 3.32 0.389 62 20116.68 167 16.7 15.6 13.84 531 760 76 71 63 4.55 3.32 0.389 63 12 91.8158.4 21.3 18.3 21.6 306 528 71 61 72 3.33 2.38 0.391 64 12 104.7 169.215 10.8 30.6 349 564 50 36 102 3.33 2.62 0.401 65 12 122.15 197.4 17.512.6 35.7 349 564 50 36 102 2.86 2.62 0.401 66 20 125 161 16.1 12.310.73 559 720 72 55 48 4.47 4.47 0.401 67 20 125 161 16.1 12.3 10.73 559720 72 55 48 4.47 4.47 0.401 68 12 117 180 24.5 21 20.5 234 360 49 42 412 2.86 0.409 69 12 122.77 187.7 23.7 10.9 29.39 259 396 50 23 62 2.112.89 0.409 70 12 129.85 193.2 23.8 13.3 28.35 371 552 68 38 81 2.86 3.050.41 71 12 111.3 165.6 20.4 11.4 24.3 371 552 68 38 81 3.33 3.05 0.41 7212 103.6 158.4 20 9.2 24.8 259 396 50 23 62 2.5 2.89 0.41 73 12 129.5198 25 11.5 31 259 396 50 23 62 2 2.89 0.41 74 12 110.33 168.7 21.3 9.826.41 259 396 50 23 62 2.35 2.89 0.411 75 20 154.5 199 19.9 15.2 13.27559 720 72 55 48 3.62 4.47 0.411 76 20 154.5 199 19.9 15.2 13.27 559 72072 55 48 3.62 4.47 0.411 77 24 133.5 165.6 12.9 8.7 11.4 445 552 43 2938 3.33 5.16 0.414 78 24 133.5 165.6 12.9 8.7 11.4 445 552 43 29 38 3.335.16 0.414 79 12 86.1 162 21 19.8 24.6 287 540 70 66 82 3.33 2.13 0.41480 12 100.45 189 24.5 23.1 28.7 287 540 70 66 82 2.86 2.13 0.414 81 12131.2 198.4 24 17.6 27.21 246 372 45 33 51 1.88 2.95 0.415 82 12 114.8173.6 21 15.4 23.8 246 372 45 33 51 2.14 2.95 0.415 83 12 123 186 22.516.5 25.5 246 372 45 33 51 2 2.95 0.415 84 12 106.6 161.2 19.5 14.3 22.1246 372 45 33 51 2.31 2.95 0.415 85 12 105.2 177.6 20.4 18.8 26.4 263444 51 47 66 2.5 2.45 0.417 86 12 130 187.2 27.6 11.2 25.6 325 468 69 2864 2.5 3.27 0.419 87 20 137.5 185 20 15 13 550 740 80 60 52 4 3.89 0.41988 20 137.5 185 20 15 13 550 740 80 60 52 4 3.89 0.419 89 20 129.25173.9 18.8 14.1 12.22 550 740 80 60 52 4.26 3.89 0.419 90 20 129.25173.9 18.8 14.1 12.22 550 740 80 60 52 4.26 3.89 0.419 91 20 134.75181.3 19.6 14.7 12.74 550 740 80 60 52 4.08 3.89 0.419 92 20 134.75181.3 19.6 14.7 12.74 550 740 80 60 52 4.08 3.89 0.419 93 20 126.5 170.218.4 13.8 11.96 550 740 80 60 52 4.35 3.89 0.419 94 20 126.5 170.2 18.413.8 11.96 550 740 80 60 52 4.35 3.89 0.419 95 20 148.5 199.8 21.6 16.214.04 550 740 80 60 52 3.7 3.89 0.419 96 20 148.5 199.8 21.6 16.2 14.04550 740 80 60 52 3.7 3.89 0.419 97 20 123.75 166.5 18 13.5 11.7 550 74080 60 52 4.44 3.89 0.419 98 20 123.75 166.5 18 13.5 11.7 550 740 80 6052 4.44 3.89 0.419 99 20 143 192.4 20.8 15.6 13.52 550 740 80 60 52 3.853.89 0.419 100 20 143 192.4 20.8 15.6 13.52 550 740 80 60 52 3.85 3.890.419

These parameters for the gearbox, where the radially inner planets areall the same size and the radially outer planets are all the same size,but not necessarily the same as the inner, are believed to work:

TABLE 7 Diameter Teeth # Outer Planet Planet Outer Planet Planet TotalSolution Planets Sun Ring Outer Inner Sun Ring Outer Inner Pitch RatioError 1 25 165.00 195.00 9.90 13.80 550 650 33 46 3.33 6.50 0.028 2 22137.18 184.90 12.20 17.35 506 682 45 64 3.69 3.87 0.028 3 18 115.20172.80 18.40 15.73 432 648 69 59 3.75 3.00 0.030 4 18 129.60 194.4020.70 17.70 432 648 69 59 3.33 3.00 0.030 5 25 144.00 180.00 8.40 15.60600 750 35 65 4.17 5.00 0.031 6 25 156.00 195.00 9.10 16.90 600 750 3565 3.85 5.00 0.031 7 22 115.82 156.10 10.30 14.65 506 682 45 64 4.373.88 0.036 8 22 110.00 167.20 15.80 16.20 550 836 79 81 5.00 2.92 0.0389 22 118.35 179.90 17.00 17.43 550 836 79 81 4.65 2.92 0.040 10 19142.50 180.50 13.00 15.50 570 722 52 62 4.00 4.75 0.056 11 22 118.80167.20 9.20 19.60 594 836 46 98 5.00 3.45 0.058 12 39 129.62 168.5011.30 9.97 390 507 34 30 3.01 4.33 0.063 13 39 143.38 186.40 12.50 11.03390 507 34 30 2.72 4.33 0.066 14 38 159.60 199.50 9.80 12.95 456 570 2837 2.86 5.00 0.068 15 38 136.80 171.00 8.40 11.10 456 570 28 37 3.335.00 0.068 16 42 123.20 156.80 10.40 8.00 462 588 39 30 3.75 4.67 0.07217 42 138.60 176.40 11.70 9.00 462 588 39 30 3.33 4.67 0.072 18 42154.00 196.00 13.00 10.00 462 588 39 30 3.00 4.67 0.072 19 34 142.80163.20 9.00 6.00 476 544 30 20 3.33 8.00 0.073 20 34 166.60 190.40 10.507.00 476 544 30 20 2.86 8.00 0.073 21 32 136.00 184.00 15.00 11.50 544736 60 46 4.00 3.83 0.081 22 32 122.40 165.60 13.50 10.35 544 736 60 464.44 3.83 0.081 23 17 132.60 198.90 15.60 24.97 170 255 20 32 1.28 3.000.083 24 17 129.20 193.80 15.20 24.33 170 255 20 32 1.32 3.00 0.083 2517 115.60 173.40 13.60 21.77 170 255 20 32 1.47 3.00 0.083 26 17 119.00178.50 14.00 22.41 170 255 20 32 1.43 3.00 0.083 27 17 122.40 183.6014.40 23.05 170 255 20 32 1.39 3.00 0.083 28 17 125.80 188.70 14.8023.69 170 255 20 32 1.35 3.00 0.083 29 17 108.80 163.20 12.80 20.49 170255 20 32 1.56 3.00 0.083 30 17 112.20 168.30 13.20 21.13 170 255 20 321.52 3.00 0.083 31 17 105.40 158.10 12.40 19.85 170 255 20 32 1.61 3.000.083 32 30 132.00 180.00 12.40 14.40 330 450 31 36 2.50 3.75 0.086 3330 129.87 177.10 12.20 14.17 330 450 31 36 2.54 3.75 0.088 34 22 130.90161.70 11.90 10.50 374 462 34 30 2.86 5.25 0.089 35 22 149.60 184.8013.60 12.00 374 462 34 30 2.50 5.25 0.089 36 18 132.30 182.70 10.5022.75 378 522 30 65 2.86 3.63 0.089 37 18 113.40 156.60 9.00 19.50 378522 30 65 3.33 3.63 0.089 38 18 119.70 165.30 9.50 20.58 378 522 30 653.16 3.63 0.089 39 18 138.60 191.40 11.00 23.83 378 522 30 65 2.73 3.630.089 40 18 126.00 174.00 10.00 21.66 378 522 30 65 3.00 3.63 0.089 4130 134.13 182.90 12.60 14.63 330 450 31 36 2.46 3.75 0.089 42 19 117.93179.70 13.30 22.76 399 608 45 77 3.38 2.91 0.096 43 23 138.00 172.5015.90 7.50 460 575 53 25 3.33 5.00 0.097 44 23 155.36 194.20 17.90 8.45460 575 53 25 2.96 5.00 0.098 45 17 107.93 166.80 20.20 14.42 187 289 3525 1.73 2.83 0.184 46 17 126.63 195.70 23.70 16.92 187 289 35 25 1.482.83 0.185 47 22 126.50 187.00 14.50 20.00 506 748 58 80 4.00 3.09 0.10548 22 143.00 176.00 15.50 8.50 286 352 31 17 2.00 5.33 0.107 49 22158.68 195.30 17.20 9.43 286 352 31 17 1.80 5.33 0.109 50 22 155.92191.90 16.90 9.27 286 352 31 17 1.83 5.33 0.109 51 22 130.08 160.1014.10 7.73 286 352 31 17 2.20 5.33 0.109 52 22 127.32 156.70 13.80 7.57286 352 31 17 2.25 5.33 0.109 53 25 137.20 176.40 9.80 14.90 350 450 2538 2.55 4.50 0.112 54 25 140.00 180.00 10.00 15.20 350 450 25 38 2.504.50 0.112 55 25 138.60 178.20 9.90 15.05 350 450 25 38 2.53 4.50 0.11256 25 135.80 174.60 9.70 14.75 350 450 25 38 2.58 4.50 0.112 57 25134.40 172.80 9.60 14.59 350 450 25 38 2.60 4.50 0.112 58 25 154.00198.00 11.00 16.72 350 450 25 38 2.27 4.50 0.112 59 25 155.40 199.8011.10 16.87 350 450 25 38 2.25 4.50 0.112 60 25 126.00 162.00 9.00 13.68350 450 25 38 2.78 4.50 0.112 61 25 148.40 190.80 10.60 16.11 350 450 2538 2.36 4.50 0.112 62 25 151.20 194.40 10.80 16.42 350 450 25 38 2.314.50 0.112 63 25 152.60 196.20 10.90 16.57 350 450 25 38 2.29 4.50 0.11264 25 149.80 192.60 10.70 16.27 350 450 25 38 2.34 4.50 0.112 65 25133.00 171.00 9.50 14.44 350 450 25 38 2.63 4.50 0.112 66 25 124.60160.20 8.90 13.53 350 450 25 38 2.81 4.50 0.112 67 25 123.20 158.40 8.8013.38 350 450 25 38 2.84 4.50 0.112 68 25 121.80 156.60 8.70 13.23 350450 25 38 2.87 4.50 0.112 69 25 127.40 163.80 9.10 13.83 350 450 25 382.75 4.50 0.112 70 25 128.80 165.60 9.20 13.99 350 450 25 38 2.72 4.500.112 71 25 131.60 169.20 9.40 14.29 350 450 25 38 2.66 4.50 0.112 72 25130.20 167.40 9.30 14.14 350 450 25 38 2.69 4.50 0.112 73 25 141.40181.80 10.10 15.35 350 450 25 38 2.48 4.50 0.112 74 25 142.80 183.6010.20 15.51 350 450 25 38 2.45 4.50 0.112 75 25 144.20 185.40 10.3015.66 350 450 25 38 2.43 4.50 0.112 76 25 145.60 187.20 10.40 15.81 350450 25 38 2.40 4.50 0.112 77 25 147.00 189.00 10.50 15.96 350 450 25 382.38 4.50 0.112 78 25 142.50 187.50 12.90 14.40 475 625 43 48 3.33 4.170.113 79 25 125.93 165.70 11.40 12.73 475 625 43 48 3.77 4.17 0.115 8021 130.13 185.90 18.00 14.76 441 630 61 50 3.39 3.33 0.116 81 19 140.18185.40 11.90 19.04 589 779 50 80 4.20 4.10 0.119 82 35 175.00 192.508.50 8.00 350 385 17 16 2.00 11.00 0.119 83 34 150.09 180.10 10.30 8.54510 612 35 29 3.40 6.00 0.121 84 19 117.80 155.80 10.00 16.00 589 779 5080 5.00 4.10 0.122 85 20 117.59 170.30 14.80 16.42 580 840 73 81 4.933.23 0.124 86 34 153.00 183.60 10.50 8.70 510 612 35 29 3.33 6.00 0.12487 20 116.00 168.00 14.60 16.20 580 840 73 81 5.00 3.23 0.125 88 32172.80 192.00 8.40 9.60 576 640 28 32 3.33 10.00 0.125 89 23 138.67184.90 10.30 18.34 552 736 41 73 3.98 4.00 0.126 90 24 116.72 158.4013.20 11.46 336 456 38 33 2.88 3.80 0.127 91 24 124.67 169.20 14.1012.24 336 456 38 33 2.70 3.80 0.127 92 24 125.56 170.40 14.20 12.33 336456 38 33 2.68 3.80 0.127 93 24 126.44 171.60 14.30 12.42 336 456 38 332.66 3.80 0.127 94 24 120.25 163.20 13.60 11.81 336 456 38 33 2.79 3.800.127 95 24 122.91 166.80 13.90 12.07 336 456 38 33 2.73 3.80 0.127 9624 117.60 159.60 13.30 11.55 336 456 38 33 2.86 3.80 0.127 97 24 115.83157.20 13.10 11.38 336 456 38 33 2.90 3.80 0.127 98 24 114.95 156.0013.00 11.29 336 456 38 33 2.92 3.80 0.127 99 24 123.79 168.00 14.0012.16 336 456 38 33 2.71 3.80 0.127 100 24 121.14 164.40 13.70 11.90 336456 38 33 2.77 3.80 0.127

Whether by material choice, stiffness reducing geometric features, orboth, the planet gear should have a torsional stiffness such that theinner and outer gears may flex torsionally enough to take up anymanufacturing tolerance in the gears and ensure proper gear toothcontact on both the inner and outer gearsets, while retaining enoughtorsional stiffness to keep the inner and outer planet gears axiallyaligned in the system and able to transmit a local torque of a magnitudecorrelating to the intended maximum torque of the gear system. Inaddition, the bending stiffness of the planet gear should be sufficientto prevent slipping of the gear teeth due to planet bending deflection.

In the claims, the word “comprising” is used in its inclusive sense anddoes not exclude other elements being present. The indefinite articles“a” and “an” before a claim feature do not exclude more than one of thefeature being present. Each one of the individual features describedhere may be used in one or more embodiments and is not, by virtue onlyof being described here, to be construed as essential to all embodimentsas defined by the claims.

The various embodiments described above can be combined to providefurther embodiments. These and other changes can be made to theembodiments in light of the above-detailed description. In general, inthe following claims, the terms used should not be construed to limitthe claims to the specific embodiments disclosed in the specificationand the claims, but should be construed to include all possibleembodiments along with the full scope of equivalents to which suchclaims are entitled. Accordingly, the claims are not limited by thedisclosure.

1. A torque transfer device, comprising: plural planets arranged forplanetary rotation about one or more sun gears and within one or morering gears, the plural planets each including a respective firstplanetary gear set comprising plural planetary gears connected to rotatetogether and having different pitch diameters; a first output gear ofthe one or more sun gears or one or more ring gears being arranged tomesh with a respective planetary gear of each first planetary gear set;a first reference gear of the one or more sun gears or one or more ringgears being arranged to mesh with a respective planetary gear of eachfirst planetary gear set; in which one or more of D or E or F or G,where D is the plural planets number at least 5, 6, 7, 8, 9, 10, 11, 12,13, 14, 15, 16, 17, 18, 19, or 20 and one or more of the one or more sungears and/or one or more of the one or more ring gears are formed of afirst material with yield strength-to-stiffness ratio greater than 0.10;E is the plural planetary gears each have a number of teeth on each partof different pitch diameter and the number of teeth is the same on eachplanetary gear and on each part of different pitch diameter; F is thetorque transfer device further comprises a respective first input gearof the one or more sun gears or one or more ring gears, the first inputgear being arranged to mesh with a respective planetary gear of eachfirst planetary gear set.the planetary gears are mounted on doublebearings that are each adjustable by one or more shims to adjustpre-load of the bearings; and G is the planetary gears of each planetarygear set, the sun gear and the ring gear each have one or both of aconical taper and a profile shift.
 2. The torque transfer device ofclaim 1 in which each planet further comprises a second planetary gearset corresponding to and arranged axially symmetrically with respect tothe first planetary gear set, the second planetary gear sets arranged tomesh with a corresponding second output gear of the one or more sungears or one or more ring gears and a corresponding second referencegear of the one or more sun gears or one or more ring gears.
 3. Thetorque transfer device of claim 2 in which the second planetary gear setof each planet has a gear tooth profile axially symmetric with respectto a gear tooth profile of the corresponding first planetary gear set.4. The torque transfer device of claim 2 in which the first and secondoutput gears are connected via a shim for adjusting the relative axialpositioning of the first and second output gears.
 5. The torque transferdevice of claim 2 in which the first and second reference gears areconnected via a shim for adjusting the relative axial positioning of thefirst and second reference gears.
 6. The torque transfer device claim 2in which the first reference gear and first output gear are connectedvia bearings, the bearings connected to at least one of the firstreference gear and first output gear via a shim.
 7. The torque transferdevice of claim 6 in which the bearings are connected to the at leastone of the first reference gear and first output gear via plural shimsconnected to different bearing races.
 8. The torque transfer device ofclaim 2 in which the second output gear is the first output gear or thesecond reference gear is the first reference gear or both.
 9. The torquetransfer device of claim 2 in which the first reference gear and secondreference gear are ring gears, and the first reference gear is connectedto the second reference gear via a housing portion extending through acenter hole defined by the one or more sun gears. 10-20. (canceled) 21.The torque transfer device of claim 1 in which the plural planetarygears of the first planetary gear set have the same number of teeth andcorresponding teeth of the plural planetary gears are circumferentiallyaligned.
 22. The torque transfer device of claim 21 in which the teethof the plural planetary gears of the first planetary gear set areconnected by a continuous tooth profile fill between correspondingteeth. 23-49. (canceled)
 50. The torque transfer device of claim 1further comprising a respective first input gear of the one or more sungears or one or more ring gears, the first input gear being arranged tomesh with a respective planetary gear of each first planetary gear set.51. The torque transfer device of claim 50 in which the first input gearis connected to an input member, and the first reference gear isconnected to a housing member, the input member rotatably connected tothe housing member through one or more intermediate members, the inputmember rotatably connected to an intermediate member of the one or moreintermediate members through a first set of bearings and the outputmember rotatably connected to the intermediate member or anotherintermediate member of the one or more intermediate members through asecond set of bearings.
 52. The torque transfer device of claim 50 inwhich two of the first input gear, first reference gear, and firstoutput gear are ring gears and one of the first input gear, firstreference gear, and first output gear is a sun gear.
 53. The torquetransfer device of claim 52 in which the input gear is a sun gear.54-71. (canceled)
 72. The torque transfer device of claim 1 in which Fis present and the double bearings are angular contract bearings toallow independent pre-load adjustment of the double bearings and theplural planetary gears.
 73. The torque transfer device of claim 1 inwhich F is present and the double bearings are provided in pairs. 74.The torque transfer device of claim 1 in which F is present and theplural planetary gears are each tapered gears to allow backlashadjustment.
 75. The torque transfer device of claim 74 in which thebacklash is adjusted by the one or more shims.
 76. The torque transferdevice of claim 75 in which the one or more shims are plural shimsconfigured to adjust the preload of the bearings independently of thebacklash.
 77. The torque transfer device of claim 50 in which the firstinput gear is supported by the planets.
 78. The torque transfer deviceof claim 77 in which the first input gear comprises a rotor of anelectric motor.
 79. The torque transfer device of claim 77 or claim 78in which the first input gear is a sun gear.
 80. The torque transferdevice of claim 1 in which G is present and the torque transfer devicecomprises a shim for adjusting a separation between the first outputgear and a second output gear.
 81. The torque transfer device of claim 1in which G is present and the torque transfer device comprises a shimfor adjusting a separation between the first output gear and the firstreference gear.
 82. The torque transfer device of claim 1 in which G ispresent and the torque transfer device comprises a shim for adjusting aseparation between two sun gears of the one or more sun gears.
 83. Thetorque transfer device of claim 50 in which G is present and the torquetransfer device comprises a shim for adjusting a separation between thefirst input gear and a second input gear.
 84. The torque transfer deviceof claim 1 in which G is present and both the conical taper and profileshift are present.
 85. The torque transfer device of claim 1 in which Gis present and the profile shift comprises a tapered profile shift. 86.The torque transfer device of claim 1 in which G is present and theprofile shift comprises a tapering of the gear teeth.